Premixed charge compression ignition engine with optimal combustion control

ABSTRACT

A premixed charge compression ignition engine, and a control system, is provided which effectively initiates combustion by compression ignition and maintains stable combustion while achieving extremely low nitrous oxide emissions, good overall efficiency and acceptable combustion noise and cylinder pressures. The present engine and control system effectively controls the combustion history, that is, the time at which combustion occurs, the rate of combustion, the duration of combustion and/or the completeness of combustion, by controlling the operation of certain control variables providing temperature control, pressure control, control of the mixture&#39;s autoignition properties and equivalence ratio control. The combustion control system provides active feedback control of the combustion event and includes a sensor, e.g. pressure sensor, for detecting an engine operating condition indicative of the combustion history, e.g. the start of combustion, and generating an associated engine operating condition signal. A processor receives the signal and generates control signals based on the engine operating condition signal for controlling various engine components to control the temperature, pressure, equivalence ratio and/or autoignition properties so as to variably control the combustion history of future combustion events to achieve stable, low emission combustion in each cylinder and combustion balancing between the cylinders.

[0001] This is a divisional application of pending application Ser. No.09/456,382, filed Dec. 8, 1999, which is a divisional of applicationSer. No. 08/916,437, filed Aug. 22, 1997.

TECHNICAL FIELD

[0002] This invention relates generally to a compression ignition enginearranged to internally burn a premixed charge of fuel and air usingautoignition to achieve reduced emissions while maintaining the desiredfuel economy.

BACKGROUND OF THE INVENTION

[0003] For well over 75 years the internal combustion engine has beenmankind's primary source of motive power. It would be difficult tooverstate its importance or the engineering effort expended in seekingits perfection. So mature and well understood is the art of internalcombustion engine design that most so called “new” engine designs aremerely designs made up of choices among a variety of known alternatives.For example, an improved output torque curve can easily be achieved bysacrificing engine fuel economy. Emissions abatement or improvedreliability can also be achieved with an increase in cost. Still otherobjectives can be achieved such as increased power and reduced sizeand/or weight but normally at a sacrifice of both fuel efficiency andlow cost.

[0004] The challenge to contemporary designers has been significantlyincreased by the need to respond to governmentally mandated emissionsabatement standards while maintaining or improving fuel efficiency. Inview of the mature nature of engine design, it is extremely difficult toextract both improved engine performance and emissions abatement fromfurther innovations of the basic engine designs commercially availabletoday. Yet the need for such innovations has never been greater in viewof the series of escalating emissions standards mandated for the futureby the United States government and other countries. Attempts to meetthese standards includes some designers looking for a completely newengine design.

[0005] Traditionally, there have been two primary forms of reciprocatingpiston or rotary internal combustion engines: diesel and spark ignitionengines. While these engine types have similar architecture andmechanical workings, each has distinct operating properties which arevastly different from each other. Diesel and spark ignited engineseffectively control the start of combustion (SOC) using simple, yetdistinct means. The diesel engine controls the SOC by the timing of fuelinjection. In a spark ignited engine, the SOC is controlled by the sparktiming. As a result, there are important differences in the advantagesand disadvantages of diesel and spark-ignited engines. The majoradvantage that a spark-ignited natural gas, or gasoline, engine has overa diesel engine is the ability to achieve extremely low NOx andparticulate emissions levels. The major advantage that diesel engineshave over premixed charge spark ignited engines (such as passenger cargasoline engines and lean burn natural gas engines) is higher thermalefficiency. One key reason for the higher efficiency of diesel enginesis the ability to use higher compression ratios than premixed chargespark ignited engines (the compression ratio in premixed charge sparkignited engines has to be kept relatively low to avoid knock). A secondkey reason for the higher efficiency of diesel engines lies in theability to control the diesel engine's power output without a throttle.This eliminates the throttling losses of premixed charge spark ignitedengines and results in significantly higher efficiency at part load fordiesel engines. Typical diesel engines, however, cannot achieve the verylow NOx and particulate emissions levels which are possible withpremixed charge spark ignited engines. Due to the mixing controllednature of diesel combustion a large fraction of the fuel exists at avery fuel rich equivalence ratio which is known to lead to particulateemissions. Premixed charge spark ignited engines, on the other hand,have nearly homogeneous air fuel mixtures which tend to be either leanor close to stoichiometric, resulting in very low particulate emissions.A second consideration is that the mixing controlled combustion indiesel engines occurs when the fuel and air exist at a nearstoichiometric equivalence ratio which leads to high temperatures. Thehigh temperatures, in turn, cause high NOx emissions. Lean burn premixedcharge spark ignited engines, on the other hand, burn their fuel at muchleaner equivalence ratios which results in significantly lowertemperatures leading to much lower NOx emissions. Stoichiometricpremixed charge spark ignited engines, on the other hand, have high NOxemissions due to the high flame temperatures resulting fromstoichiometric combustion. However, the virtually oxygen free exhaustallows the NOx emissions to be reduced to very low levels with athree-way catalyst.

[0006] Relatively recently, some engine designers have directed theirefforts to another type of engine which utilizes premixed chargecompression ignition (PCCI) or homogeneous charge compression ignition(HCCI), hereinafter collectively referred to as PCCI. Engines operatingon PCCI principles rely on autoignition of a relatively well premixedfuel/air mixture to initiate combustion. Importantly, the fuel and airare mixed, in the intake port or the cylinder, long before ignitionoccurs. The extent of the mixture may be varied depending on thecombustion characteristics desired. Some engines are designed and/oroperated to ensure the fuel and air are mixed into a homogeneous, ornearly homogeneous, state. Also, an engine may be specifically designedand/or operated to create a somewhat less homogeneous charge having asmall degree of stratification. In both instances, the mixture exists ina premixed state well before ignition occurs and is compressed until themixture autoignites. Importantly, PCCI combustion is characterized inthat: 1) the vast majority of the fuel is sufficiently premixed with theair to form a combustible mixture throughout the charge by the time ofignition and throughout combustion; and 2) combustion is initiated bycompression ignition. Unlike a diesel engine, the timing of the fueldelivery, for example the timing of injection, in a PCCI engine does notstrongly affect the timing of ignition. The early delivery of fuel in aPCCI engine results in a premixed charge which is very well mixed, andpreferably nearly homogeneous, thus reducing emissions, unlike thestratified charge combustion of a diesel which generates higheremissions. Preferably, PCCI combustion is characterized in that most ofthe mixture is significantly leaner than stoichiometric toadvantageously reduce emissions, unlike the typical diesel engine cyclein which a large portion, or all, of the mixture exists in a rich stateduring combustion.

[0007] An engine operating on PCCI combustion principles has thepotential for providing the excellent fuel economy of the diesel enginewhile providing NOx and particulate emissions levels that are much lowerthan that of current spark-ignited or diesel engine. For example, U.S.Pat. No. 4,768,481 to Wood discloses a process and engine that isintended to use a homogeneous mixture of fuel and air which isspontaneously ignited. A controlled rate of combustion is said to beobtained by adding exhaust products to the air-fuel mixture. Acombustion chamber is connected to the engine cylinder and fuel gas issupplied to the chamber via a check valve. A glow plug is positionedbetween the combustion chamber and the cylinder. The mixture enteringthe combustion is heated by the glow plug and by the hot walls of thecombustion chamber. The mixture ignites due to the increase intemperature and the increase in pressure resulting from compression. TheWood patent is specifically directed to a two-stroke engine, butgenerally mentions that the technology could be applied to a four-strokeengine. However, this reference fails to discuss how the exhaust gasrecirculation and glow plug would be controlled to optimize the start ofcombustion and to maintain the optimal start, and duration, ofcombustion, as load and ambient conditions change. A practicalembodiment of this engine is unlikely to be capable of effectivelycontrolling and maintaining PCCI combustion without additional controls.

[0008] U.S. Pat. No. 5,535,716 issued to Sato et al., discloses acompression ignition type engine which greatly reduces NOx emissions byintroducing an evaporated fuel/air mixture into the combustion chamberduring the intake event and early in the compression event forself-ignited combustion later in the compression event. The amount ofNOx emissions produced by this engine is about one-thirtieth of thatproduced by a diesel engine. These principles are also set forth in SAETechnical Paper No. 960081, Aoyama, T. et al., “An Experimental Study onPremixed-Charge Compression Ignition Gasoline Engine”, Feb. 26, 1996.However, these references do not specifically discuss controlling thetiming of the start of combustion and the rate of combustion. Moreover,the engine disclosed in these references only uses the heat generated bycompression to ignite the charge, without the use of any preheating.Also, these references do not suggest the controls, nor the manner ofoperating the controls, necessary to maintain stable combustion. Also,these references only disclose the use of gasoline.

[0009] U.S. Pat. No. 5,467,757 issued to Yanagihara et al., discloses adirect injection compression-ignition type engine in which fuel isinjected into a combustion chamber during the intake stroke orcompression stroke, before 60 degrees BTDC of the compression stroke, soas to reduce the amount of soot and NOx generated to substantially zero.These advantages are achieved by considerably enlarging the meanparticle size of the injected fuel from the mean particle size used inconventional combustion processes to prevent the early vaporization ofinjected fuel after injection and by making the injection timingconsiderably earlier than conventional injection timing to ensure auniformed fusion of the injected fuel in the combustion chamber.However, this reference nowhere suggests a manner of activelycontrolling the combustion history, such as the timing of the start ofcombustion and/or the duration of combustion.

[0010] Researchers have used various other names to refer to PCCIcombustion. For example, Onishi, et al. (SAE Technical Paper No. 790501,Feb. 26-Mar. 2, 1979) called it “ATAC”, which stands for “ActiveThermo-Atmosphere Combustion.” Noguchi, et al. (SAE Technical Paper No.790840, Sep. 10-13, 1979) called it “TS”, which stands for“Toyota-Soken”, and Najt, et al. (SAE Paper No. 830264, 1983) called it“CIHC”, which stands for “compression-ignited homogeneous charge.”

[0011] Onishi, et al., worked with two-stroke engines. They found thatPCCI combustion (ATAC) could be made to occur in a two-stroke engine atlow load over a wide speed range. Combustion stability was much betterthan in the standard engine and there were significant improvements infuel economy and exhaust emissions. Schlieren photography of thecombustion was carried out with results quite similar to those obtainedin their combustion studies. It was found that combustion was initiatedat many points in the combustion chamber. However, there were small timedifferences between the start of combustion of these many points. Also,the combustion reactions were found to require a relatively long timecompared to conventional spark-ignited flame propagation. To attain PCCIcombustion, the following conditions were found to be important. Thequantity of mixture and the air/fuel ratio supplied to the cylinder mustbe uniform from cycle to cycle. The scavenging “directivity” andvelocity must have cyclic regularity to ensure the correct condition ofthe residual gases remaining in the cylinder. The temperature of thecombustion chamber walls must be suitable. The scavenging passage inletmust be located at the bottom of the crankcase. It was found that atvery light loads, PCCI was not successful because charge temperatureswere too low. At very high loads, PCCI was not successful because theresidual gas quantity was too low. In between these regions, PCCIcombustion was successful.

[0012] Noguchi also obtained PCCI combustion in a two-stroke engine.Very stable combustion was observed, with low emissions of hydrocarbons(HC) and improved fuel consumption. Operation in PCCI mode was possiblebetween 800 and 3200 rpm and air/fuel ratios between 11 and 22. Deliveryratios of up to 0.5 could be achieved at idle conditions. They observedthat combustion could start at lower temperatures and pressures thanthose required for conventional diesel combustion. The combustionbehavior was different from that of conventional spark-ignitedcombustion. Ignition occurred at numerous points around the center ofthe combustion chamber and the flame spread rapidly in all directions.The combustion duration was shorter than that of conventionalcombustion. It was proven that ignition kernels were not generated fromcontaminants deposited on the combustion chamber walls (generallypresumed to be the cause of “run-on” phenomena in conventional gasolineengines). To gain a better understanding of the combustion, they set upan experimental apparatus for detecting radicals in the combustionchamber. It was found that the radicals showed higher peaks of luminousintensity that disappeared at an earlier time than with conventionalspark-ignited combustion. In the case of conventional spark-ignitioncombustion, all the radicals such as OH, CH, C₂, H, and CHO, HO₂, O wereobserved at almost the same crank angle. However, with PCCI combustion,CHO, HO₂ and O radicals were detected first, followed by HC, C₂, and Hradicals, and finally the OH radical.

[0013] Najt, et al. were able to achieve PCCI combustion in afour-stroke engine. They used a CFR single-cylinder engine with ashrouded intake valve. Several compression ratios were tried, and it wasfound that, although higher ratios would allow combustion at lowercharge gas temperatures, they also resulted in excessively fast heatrelease rates. While a compression ratio of 7.5:1 was satisfactory, acompression ratio of 10:1 was not. Intake temperatures were in the rangeof 480° K to 800° K. Their average energy release rates wereconsiderably higher than those measured by Onishi and Noguchi.

[0014] SAE Paper No. 960742, entitled “Improving the Exhaust Emissionsof Two-Stroke Engines by Applying the Activated Radical Combustion”,Ishibashi, Y. et al., 1996, is noted as disclosing yet another study ofPCCI combustion in a two-stroke engine.

[0015] Although Onishi et al., Noguchi et al., Najt et al. andIshibashi, et al. have made significant progress in understanding PCCIcombustion, these references fail to suggest a practical PCCI enginehaving a control system capable of maintaining stable, efficient PCCIcombustion with low emissions by controlling the time at whichcombustion occurs, the duration of combustion, the rate of combustionand/or the completeness of combustion. Specifically, these references donot suggest a PCCI engine and control system capable of effectivelycontrolling the start of combustion. Moreover, these references do notsuggest a system capable of actively enhancing the engine startabilityand achieving combustion balancing between the cylinders in amulti-cylinder engine.

[0016] SAE Technical Paper No. 892068, entitled “Homogeneous-ChargeCompression Ignition (HCCI) Engines”, Thring, R., Sep. 25, 1989,investigated PCCI operation of a four stroke engine. The paper foundthat PCCI required high exhaust gas recirculation (EGR) rates and highintake temperatures. It was shown that PCCI combustion produces fueleconomy results comparable to a direct injection diesel engine and, thatunder favorable conditions, i.e. equivalence ratio of 0.5 and EGR rateof 23%, produces very low cyclic irregularity. This study also concludedthat before PCCI can be made practical, it will be necessary to operatean engine in the PCCI mode without the need to supply large amounts ofheat energy to the intake. The paper suggests two possibilities: the useof heated surfaces in the combustion chamber and the use of multi-stageturbocharging without intercoolers. However, although this papersuggests further investigating the effects of EGR and intake temperatureon the timing of the start of combustion, this paper fails to disclose asystem for effectively achieving active control of the start andduration of combustion.

[0017] U.S. Pat. No. 5,476,072 to Inventor discloses another example ofa PCCI engine which includes a cylinder head design that preventsexcessive stresses and structural damage that PCCI engines inherentlytend to cause. Specifically, the head includes a movable accumulatorpiston which moves to limit the peak cylinder pressure and temperature.However, control over the movement of the piston is merely passive and,therefore, this engine is unlikely to effectively stabilize combustion.Moreover, this reference nowhere suggests controlling the timing atwhich rapid combustion occurs, nor how such control could beaccomplished.

[0018] An October 1951 publication entitled “Operating directions—LOHMANN BICYCLE MOTOR” discloses a two-stroke engine operating on PCCIcombustion principles. Compression ratio is continuously adjustablebased on outside temperature, fuel, speed and load. However, this enginerequires the operator control the compression ratio manually. Therefore,this engine could not provide effective active control of combustion toensure efficient combustion with low emissions throughout all operatingconditions. Also, manual adjustment of compression ratio alone, withoutautomatic temperature, equivalence ratio and/or autoignition propertycontrol, will not result in stable, optimized combustion throughout alloperating conditions.

[0019] Conventional “dual fuel” engines operate on both a gaseous fuelmixture and diesel fuel. However, conventional dual fuel engines utilizethe timing of the injection of diesel fuel to control the SOC of thefuel/air mixture received from the intake duct. In order to achieve thisresult, dual fuel engines inject the diesel fuel at approximately topdead center. In addition, the quantity of diesel fuel injected in a dualfuel engine is sufficient to ensure that the gaseous fuel in thecombustion chamber ignites and bums virtually completely. As a result,dual fuel engines produce emissions similar to most conventional dieseland natural gas engines. In particular, in known dual fuel engines usingdiesel fuel and natural gas at high load, only a small amount of dieselfuel is required to start ignition and the emissions produced would besimilar to a spark ignited natural gas engine. Under other conditionswhen substantial diesel fuel is injected, the emissions produced wouldbe similar to a conventional diesel engine.

[0020] Consequently, there is a need for an engine operating on PCCIprinciples which includes a combustion control system capable ofeffectively controlling the timing of the start of combustion orlocation of combustion, and the rate or duration of combustion duringengine operation.

SUMMARY OF THE INVENTION

[0021] A general objective of the subject invention is to overcome thedeficiencies of the prior art by providing a practical PCCI engine and acontrol system for effectively and efficiently operating the PCCIengine.

[0022] Another object of the present invention is to provide a PCCIengine and control scheme for controlling the engine in a manner tooptimally minimize emissions, especially oxides of nitrogen andparticulate emissions, while maximizing efficiency.

[0023] Yet another object of the present invention is to provide a PCCIengine and control system for optimally controlling the combustionhistory of subsequent combustion events to effectively control thecombustion event.

[0024] Still another object of the present invention is to provide aPCCI engine and control system for effectively controlling PCCIcombustion in such a manner to achieve acceptable cylinder pressurewhile minimizing combustion noise.

[0025] A further object of the present invention is to provide a PCCIengine and control system which operates to actively control thecombustion history of future combustion events during engine operationby sensing an engine operating condition indicative of the combustionhistory.

[0026] A still further object of the present invention is to provide aPCCI engine and control system which effectively controls various engineoperating control variables to control the time at which the combustionevent occurs during the compression and expansion events of the engine.

[0027] Yet another object of the present invention is to provide a PCCIengine and control system which effectively ensures that combustionoccurs at an appropriate crank angle during the engine cycle to ensurestable combustion, low emissions, acceptable pressure levels and optimumefficiency.

[0028] Another object of the present invention is to provide a PCCIengine and control system which effectively controls the temperature,pressure, equivalence ratio and/or air/fuel mixture autoignitionproperties to precisely control the timing of the start of combustion.

[0029] A still further object of the present invention is to provide aPCCI engine and control system which effectively achieves continuous,stable PCCI combustion while achieving acceptable cylinder pressures andthe desired brake mean effective pressure.

[0030] Yet another object of the present invention is to provide a PCCIengine and control system which effectively controls the commencement ofcombustion and the combustion rate so as to ensure that substantiallyall of the combustion process occurs within an optimal crank anglelimit, i.e. 20 degrees BTDC through 35 degrees ATDC, while minimizingemissions and maximizing efficiency.

[0031] Another object of the present invention is to provide a PCCIengine which can be easily started.

[0032] Still another object of the present invention is to provide amulti-cylinder engine and control system which effectively minimizesvariations in the combustion events of the cylinders.

[0033] Yet another object of the present invention is to provide amulti-cylinder engine and control system which effectively controls thestart of combustion to achieve stable, low emission, efficientcombustion throughout exposure to changes in engine load and ambientconditions.

[0034] Another object of the present invention is to provide a controlsystem for a engine which effectively detects or senses the start ofcombustion to provide feedback control and then controls the operatingconditions of the engine to optimize the start of combustion.

[0035] Still another object of the present invention is to provide aPCCI engine and control system which effectively minimizes the unburnedhydrocarbon and carbon monoxide emissions.

[0036] The above objects and others are achieved by providing a premixedcharge compression ignition internal combustion engine, comprising anengine body, a combustion chamber formed in the engine body andcombustion history control system for controlling a combustion historyof future combustion events to reduce emissions and optimize efficiency.The combustion history control system includes at least one of atemperature control system for varying the temperature of the mixture offuel and air, a pressure control system for varying the pressure of themixture, an equivalence ratio control system for varying an equivalenceratio of the mixture and a mixture autoignition property control systemfor varying an autoignition property of the mixture. The engine furtherincludes an operating condition detecting device for detecting an engineoperating condition indicative of the combustion history and generatingan engine operating condition signal indicative of the engine operatingcondition, and a processor for receiving the engine operating conditionsignal, determining a combustion history value based on the engineoperating condition signal, and generating one or more control signalbased on the combustion history value. The one or more control signalsare used to control at least one of the temperature control system, thepressure control system, the equivalence ratio control system and themixture autoignition property control system to variably control thecombustion history of future combustion events.

[0037] The engine operating condition detecting device may include astart of combustion sensor for sensing the start of combustion andgenerating a start of combustion signal. Also, the combustion historyvalue may be determined based on the start of combustion signal. Theengine operating condition detecting device may be a cylinder pressuresensor.

BRIEF DESCRIPTION OF THE DRAWINGS

[0038]FIG. 1a is a schematic diagram of one embodiment of the presentinvention showing a single cylinder of the engine of FIG. 1b andassociated control system;

[0039]FIG. 1b is a schematic diagram of a multi-cylinder engine of thepresent invention;

[0040]FIG. 2 is a graph showing cylinder pressure and heat release rateas a function of crank angle for the PCCI engine of the presentinvention;

[0041]FIG. 3 is a graph showing the apparent heat release rate as afunction of crank angle for several different engine operatingconditions;

[0042]FIG. 4a is a graph showing knock intensity as a function of timefor a given set of operating conditions;

[0043]FIG. 4b is a graph showing gross indicated mean effective pressure(GIMEP) as a function of time;

[0044]FIG. 4c is a graph showing peak pressure as a function of time forthe same conditions of FIGS. 4a and 4 b;

[0045]FIG. 5 is a graph showing apparent heat release rate as a functionof crank angle and illustrating the increase in the heat release rateduration as the combustion or heat release location or timing isretarded;

[0046]FIG. 6 is a graph showing cylinder pressure as a function of crankangle and illustrating the decrease in peak cylinder pressure as theheat release rate retards;

[0047]FIG. 7a is a graph showing GIMEP as a function of intake manifoldtemperature for two different engine speed cases;

[0048]FIG. 7b is a graph showing the coefficient of variation of GIMEPas a function of intake manifold temperature for two different enginespeed cases;

[0049]FIG. 7c is a graph showing peak cylinder pressure as a function ofintake manifold temperature for two different engine speeds;

[0050]FIG. 7d is a graph showing the start of combustion as a functionof intake manifold temperature for two different engine speeds;

[0051]FIG. 7e is a graph showing heat release duration in crank angledegrees as a function of intake manifold temperature for two differentengine speeds;

[0052]FIG. 7f is a graph showing heat release duration in time as afunction of intake manifold temperature for two different engine speeds;

[0053]FIG. 7g is a graph showing gross indicated thermal efficiency as afunction of intake manifold temperature for two different engine speeds;

[0054]FIG. 7h is a graph showing fuel specific hydrocarbons as afunction of intake manifold temperature for two different engine speeds;

[0055]FIG. 7i is a graph showing fuel specific carbon monoxide as afunction of intake manifold temperature for two different engine speeds;

[0056]FIG. 7j is a graph showing fuel specific oxides of nitrogenemissions as a function of intake manifold temperature for two differentengine speeds;

[0057]FIG. 7k is a graph showing noise as a function of intake manifoldtemperature for two different engine speeds;

[0058]FIG. 8 is a graph showing apparent heat release rate as a functionof crank angle for three different intake manifold temperatures;

[0059]FIG. 9 is a graph showing both the start of combustion andcombustion duration as a function of wall temperature;

[0060]FIG. 10 is a graph showing both the start and end of combustion asa function of crank angle for a given time period, and GIMEP for thesame time period, wherein a glow plug is cycled;

[0061]FIG. 11 is a graph showing the apparent heat release rate as afunction of crank angle for the glow plug transient of FIG. 10;

[0062]FIG. 12 discloses one embodiment of an end cylinder compensatingsystem of the present invention for providing cylinder-to-cylindertemperature control;

[0063]FIG. 13 is a schematic diagram of a second embodiment of the endcylinder compensating device for providing cylinder-to-cylindertemperature control;

[0064]FIG. 14 is a graph showing the effects of changing intake andexhaust valve opening and closing events on top dead center (TDC)temperature;

[0065]FIG. 15 is a graph showing the effects of changing intake andexhaust valve opening and closing events, and variable compressionratio, on the residual mass fraction and temperature at top dead center;

[0066]FIG. 16 is a graph showing both cylinder pressure and heat releaseas a function of crank angle for different exhaust valve lash settings;

[0067]FIG. 17 is a graph showing the effects of varying exhaust gasrecirculation (EGR) on the location of the heat release rate relative tothe crank angle and the effect of variations in EGR on the magnitude ofthe heat release rate;

[0068]FIG. 18 is a graph showing the effect of varying the EGR rate onthe timing of the start of combustion;

[0069]FIG. 19 is a schematic of an improved engine of the presentinvention having one cylinder operating under PCCI conditions tooptimize the use of EGR;

[0070]FIG. 20 is a graph showing the effects of changing compressionratio on the temperature at top dead center;

[0071]FIG. 21 is a graph showing the start of combustion as a functionof intake manifold temperature and the effects of changing thecompression ratio on the start of combustion and intake manifoldtemperature;

[0072]FIG. 22a is a partial cross sectional view of one cylinder of thePCCI engine of the present invention including one embodiment of acompression ratio varying

[0073]FIG. 22b is a partial cross sectional view of one cylinder of thePCCI engine of the present invention showing a second embodiment of acompression ratio varying

[0074]FIG. 22c is a partial cross sectional view of one cylinder of thepresent PCCI engine showing a third embodiment of the compression ratiovarying device;

[0075]FIG. 22d is a partial cross sectional view of a single cylinder ofthe present PCCI engine showing a fourth embodiment of the compressionratio varying device of the present invention;

[0076]FIG. 23 is a schematic diagram of an opposed piston PCCI engine ofthe present invention including a variable phase shifting mechanism forvarying the compression ratio;

[0077]FIG. 24 is a side view of the differential mechanism used in thevariable phase shifting mechanism of FIG. 23;

[0078]FIG. 25 is a graph showing compression ratio as a function of thedegrees out of phase of two pistons in the opposed piston engine, forexample, of FIG. 23 illustrating various compression ratio settings;

[0079]FIG. 26 is a graph showing cylinder volume as a function of crankangle of a reference piston in an opposed piston PCCI engine which showsthat the compression ratio decreases as the pistons become more out ofphase;

[0080]FIG. 27 is a graph showing the effects of changing intake andexhaust valve opening and closing events, and varying the compressionratio, on the percent of baseline airflow rate and the TDC temperature;

[0081]FIG. 28 is a graph showing the effects of changes and intake inexhaust valve opening and closing events, and varying the compressionratio, on the diesel equivalent brake specific fuel consumption and TDCtemperature;

[0082]FIG. 29 is a graph showing the effects of changes and intake inexhaust valve opening and closing events, and variations in compressionratio, on peak cylinder pressure and TDC temperature;

[0083]FIG. 30 is a graph showing the effects of water injection onintake manifold temperature and temperature at top dead center;

[0084]FIG. 31 a is a graph showing the combustion duration in crankangle degrees as a function of intake manifold pressure (IMP);

[0085]FIG. 31b is a graph showing combustion duration in time as afunction of

[0086]FIG. 31c is a graph showing the effect of changes in IMP on themagnitude and timing or location of the heat release rate;

[0087]FIG. 31d is a graph showing the start of combustion timing andcrank angle degrees as a function of IMP;

[0088]FIG. 31e is a graph showing fuel specific hydrocarbons as afunction of IMP;

[0089]FIG. 31f is a graph showing GIMEP as a function of IMP;

[0090]FIG. 31g is a graph showing gross indicated thermal efficiency asa function of IMP;

[0091]FIG. 31h is a graph showing fuel specific carbon monoxide as afunction of IMP;

[0092]FIG. 31i is a graph showing fuel specific oxides of nitrogenemissions as a function of IMP;

[0093]FIG. 31j is a graph showing the coefficient of variation of GIMEPas a function of IMP;

[0094]FIG. 31k is a graph showing the peak cylinder pressure as afunction of IMP;

[0095]FIG. 31l is a graph showing noise as a function of IMP;

[0096]FIG. 31m is a graph showing the effects of increasing IMP on peakcylinder pressure and GIMEP;

[0097]FIG. 32 is a graph showing the effect of various trace species ona start of combustion and temperature;

[0098]FIG. 33 is a graph showing the effects of additional amounts ofozone on advancing the start of combustion;

[0099]FIG. 34 is a graph showing the effect of varying the type of fuelused in the present PCCI engine on the start of combustion wherein theincrease in temperature indicates the start of combustion;

[0100]FIG. 35 is a graph showing the apparent heat release duration as afunction of equivalence ratio;

[0101]FIG. 36 is a graph showing the start of combustion in crank angledegrees as a function of equivalence ratio;

[0102]FIG. 37 is a graph showing the effects of variations inequivalence ratio on the start of combustion wherein an increase intemperature indicates the start of combustion;

[0103]FIG. 38 is a graph showing the effects of variations in theequivalence ratio on the magnitude and timing, or location, of the heatrelease rate;

[0104]FIG. 39 is a graph showing the effects of equivalence ratio on thecompressor pressure ratio and the compressor outlet temperature;

[0105]FIG. 40 is a graph showing the effects of varying the equivalenceratio on the brake specific fuel consumption;

[0106]FIG. 41 is a graph showing the differences in pumping meaneffective pressure and GIMEP for two differently sized turbine casings;

[0107]FIG. 42 is a graph showing the diesel equivalent BSFC and BMEP fortwo differently sized turbine casings;

[0108]FIG. 43 is a graph showing the turbine rotor speed and intakemanifold pressure for two differently sized turbine casings;

[0109]FIG. 44 is a graph showing the fuel specific oxides of nitrogenemissions for PCCI combustion with various fuels in comparison to atypical compression ignition diesel engine;

[0110]FIG. 45 is a graph showing emissions as a function of enginespeed;

[0111]FIG. 46 is a graph showing emissions as a function of temperatureat bottom dead center;

[0112]FIG. 47 is a graph showing fuel specific carbon monoxide as afunction of end of combustion flame temperature;

[0113]FIGS. 48a-50 b are partial cross sectional views of a singlecylinder of the PCCI engine of the present invention showing analternative embodiment including various crevice minimizing features;and

[0114]FIG. 51 is a graph showing the effects of various percentages ofdiesel pilot injections on the heat release rate location and shape.

DETAILED DESCRIPTION OF THE INVENTION

[0115] The present invention is directed to an improved premixed chargecompression ignition (PCCI) engine and control scheme for controllingthe engine in a manner to optimally minimize emissions while maximizingefficiency. For the purposes of this application, PCCI refers to anyengine or combustion process in which: 1) the vast majority of the fuelis sufficiently premixed with the air to form a combustible mixturethroughout the charge by the time of ignition and throughout combustion;and 2) combustion is initiated by compression ignition. PCCI also refersto any compression ignition engine or combustion process in which thefuel and air are premixed long before ignition. As a result, the timingof injection of the fuel in the PCCI engine does not affect the timingof ignition of the fuel/air mixture. Also, it should be understood thatPCCI is meant to encompass homogeneous charge compression ignition(HCCI) engines and processes wherein the mixture exists in ahomogeneous, or nearly homogeneous state, at the start of combustion. Inthe present invention, the fuel/air mixture is thoroughly mixed to forma very lean homogeneous mixture, or is mixed in a manner to form a lesshomogeneous mixture with a desired air/fuel stratification, to ensurerelatively even, low flame temperatures which result in extremely lowoxides of nitrogen (NOx) emissions. It should be understood the someengines operate under PCCI conditions continuously while other enginesmay operate under PCCI conditions for only a limited period of operationeither by design or inadvertently.

[0116] Applicants have recognized that the key to producing acommercially viable PCCI engine lies in the control of the combustionhistory of subsequent or future combustion events in such a manner so asto result in extremely low NOx emissions combined with very good overallefficiency, combustion noise control and with acceptable cylinderpressure. The combustion history may include the time at whichcombustion occurs (combustion timing), the rate of combustion (heatrelease rate), the duration of combustion and/or the completeness ofcombustion. Applicants have determined that the combustion history, andespecially the combustion timing, is sensitive to, and varies dependingon, a variety of factors including changes in load and ambientconditions. The engine and control system of the present inventionoperates to actively control the combustion history of future combustionevents during engine operation to ensure the desired combustion andengine operation is maintained. In the preferred embodiment, the presentengine and control system controls the combustion timing during thecompression and expansion events of the engine.

[0117]FIGS. 1a and 1 b illustrates the PCCI engine and control system ofthe present invention, indicated generally at 10. FIG. 1a shows a singleengine cylinder 12 of the multi-cylinder reciprocating piston typeengine shown in FIG. 1b. Of course, the PCCI control system of thepresent invention could be used to control PCCI combustion in an enginehaving only a single cylinder or any number of cylinders, for example, afour, six, eight or twelve cylinder internal combustion engine. Inaddition, although the present PCCI control system is primarilydiscussed with reference to a four stroke engine, the present controlsystem could be applied to a two stroke engine. Also, the PCCI system ofthe present invention may be adapted for use on any internal combustionengine having compression, combustion and expansion events, including arotary engine and a free piston engine.

[0118] As shown in FIG. 1a, a piston 14 is reciprocally mounted in thecylinder to form a combustion chamber 13. The piston transmits forcesgenerated by a combustion event into a conventional engine drive system.Referring to FIGS. 1a and 1 b, an intake air system 23 including anintake manifold 15 supplies intake air, or an air/fuel mixture to arespective intake port 26 associated with each cylinder 12. Likewise, anexhaust gas system 27 including an exhaust manifold 17 receives exhaustgases flowing from exhaust ports 31. One or more intake valves, such anintake valve 19 and one or more exhaust valves, such as exhaust valve21, are moved between open and closed positions by a conventional valvecontrol system, or a variable valve timing system, to control the flowof intake air or air/fuel mixture into, and exhaust gases out of, thecylinder, respectively.

[0119] The PCCI system 10 includes a combustion sensor 16 for sensing ordetecting an engine operating condition indicative of the combustionhistory and generating a corresponding signal 18. In the preferredembodiment, sensor 16 permits effective combustion control capability bydetecting an engine operating condition or parameter directly relatedto, or indicative of, the time at which the combustion event occursduring the compression and/or expansion strokes, i.e. preferably thestart of combustion (SOC). For example, a cylinder pressure sensor maybe provided on any or all engine cylinders for sensing, on acycle-by-cycle basis, the SOC. In this case, the sensor 16 also providesother engine condition data, such as the combustion rate, combustionduration, combustion event or heat release location and end ofcombustion data, any one of which may be used instead of the start ofcombustion data. Any conventional means for detecting the start ofcombustion may be used, for example, by sensing the very rapid increasein the cylinder pressure. Other forms of sensors could be used includingaccelerometers, ion probes, optical diagnostics, strain gages and/orfast thermocouples in the cylinder head, liner or piston. Also, torqueor RPM sensors could be used to detect changes in engine torque and RPMassociated with each combustion event. Alternatively, or additionally,an emissions sensor could be used to detect emissions having a knowncorrelation to the completeness of combustion.

[0120] Sensor 16 provides feedback control to an electronic control unit20 (ECU). ECU 20 receives signal 18, processes the signal and determinesan actual combustion history value, i.e. start of combustion value. Theactual combustion history value is then compared to a predetermineddesired combustion history value obtained, for example, from a look-uptable. Based on the comparison of the actual combustion history value tothe desired combustion history value, ECU 20 then generates a pluralityof output signals, indicated at 22, for variably controlling respectivecomponents of the system so as to effectively ensure, in the preferredembodiment, that the SOC and completion of combustion occur between 20degrees before top dead center (BTDC) during the compression stroke and35 degrees after top dead center (ATDC) during the power stroke of thepiston thereby minimizing NOx emissions while maximizing engineefficiency. The PCCI combustion control scheme is most preferablyimplemented in software contained in ECU 20 that includes a centralprocessing unit such as a micro-controller, micro-processor, or othersuitable micro-computing unit.

[0121] As discussed herein, PCCI system 10 may include variouscomponents for optimizing the combustion event. The objectives of thepresent system, i.e. low oxides of nitrogen (NOx) emissions, highefficiency, etc, may be achieved using any one of the various controlcomponents, or any combination of the components. In particular, asshown in FIG. 1b, a compressor 24 may be provided along an intake airsystem 23 upstream of intake manifold 15 for varying the boost intakepressure. Compressor 24 may be driven by any conventional means, such asan exhaust gas driven turbine 25. A bypass circuit 33 including a wastegate valve 43 may be provided in a conventional manner. A secondcompressor or supercharger 58 may be provided upstream of compressor 24.Supercharger 58 is mechanically driven by the engine drive system. Acharge air cooler 28 may also be provided downstream of compressor 24.Also, an intake air heater 30 (such as a burner, heat exchanger or anelectric heater) may be provided, for example, after cooler 28 as shownin FIG. 1b, or alternatively, upstream of compressor 24. Also, anindividual heater 29 may be provided in the intake port 26 associatedwith each cylinder 12 to provide quicker control of the intake manifoldtemperature for each cylinder to enhance both individual cylindercombustion control and balancing of combustion between the cylinders.Compressor 24, cooler 28 and heater 30 each include control devices forvarying the effect of the particular component on thepressure/temperature of the intake air or mixture. For example, a bypassvalve or waste gate 43 could be used to regulate the amount of exhaustgas supplied from the associated exhaust system, which is connected toan exhaust duct 31, to turbine 25 thereby varying the intake pressure asdesired. Similarly, a control valve could be provided in the coolingfluid flow path supplied to cooler 28 to permit variable control of thecooling effect of cooler 28. Likewise, various types of variablecontrols could be used to vary the heating effect of heater 30. Outputsignals 22 from ECU 20 are supplied to the various control devices tocontrol compressor 24, cooler 28 and heater 30 so as to variably controlthe pressure and temperature of the intake air or mixture preferably ona cycle-by-cycle basis.

[0122] In addition, the PCCI system 10 may include a plurality of fuelsupplies 32 and 34 for supplying fuels having different autoignitionproperties (for example, different octane or methane ratings, oractivation energy levels) into the intake air flow. Fuel control valves39 and 41 are used to control the amount of each fuel supply 32, 34delivered, respectively. For example, fuel may be supplied along theintake air path between cooler 28 and air heater 30 as shown in FIG. 1b.Of course, fuel could be introduced at various locations along theintake of the engine, such as upstream of the cooler, e.g. upstream ofthe compressor. Alternatively, the fuel could be injected, by forexample an injector 35, into the respective intake duct 26 associatedwith each cylinder, as shown in FIG. 1a.

[0123] The present PCCI system 10 also importantly includes a variablecompression ratio means 38 for varying the compression ratio so as toadvantageously advance or retard the combustion event as desired. Forexample, variable compression ratio means 38 may be in the form of acontrol mechanism for varying the shape of the combustion chamber orheight of the piston to vary the effective compression ratio. Theeffective compression ratio could also be varied by varying the timingof closing of intake valve 19 as discussed more fully hereinbelow. Thevariations in the timing of opening and closing of the intake andexhaust valves may be accomplished using any conventional variable valvetiming actuator system capable of receiving signals from ECU 20 andeffectively varying the opening and/or closing of the valves inaccordance with the principles set forth hereinbelow.

[0124] In addition, in-cylinder diluent injection may be accomplishedusing an injector 40 for injecting a gas or liquid, e.g. air, nitrogen,carbon dioxide, exhaust gas, water, etc., into the cylinder to vary thetemperature and the temperature distribution in the cylinder so as tocontrol the combustion event. Similarly, a diluent may be injected intointake duct 26 using, for example, an injector 42.

[0125] The present PCCI system may also include a fuel injector 36 forinjecting fuel 37, e.g. diesel fuel, directly into the combustionchamber. Fuel 37 would be injected either early in the compressionevent, preferably approximately between 180 degrees and 60 degrees BTDC,as described below, or later in the compression event near TDC.

[0126] By injecting the fuel 37 early in the compression event, it ismuch more thoroughly mixed with the fuel/air mixture received from theintake duct than would be the case for a diesel engine, thus ensuring amore desirable combustion process, in particular the fuel will bum at aleaner equivalence ratio which results in much lower NOx emissions. Thestart or initiation of the combustion (SOC) of the fuel/air mixturereceived from the intake duct may be varied by controlling the quantityof fuel 37 injected. For instance, an earlier combustion event may beachieved by increasing the quantity of fuel 37 while the timing of thecombustion event may be delayed by decreasing the quantity of fuel 37injected.

[0127] By injecting the fuel 37 later in the compression stroke, that isnear TDC, conventional diesel fuel injection systems can be used. Thisapproach could be combined with the introduction of one or moreadditional types of fuel in the intake manifold to achieve a PCCI modeof operation. In particular, the fuel injected into the intake manifoldwould have a higher excess air ratio. The excess air ratio is the actualair-fuel ratio of the engine divided by the air-fuel ratio atstoichiometric conditions. For the very lean excess air ratio,combustion along a flame front is impossible. However, autoignition ispossible thereby allowing combustion of a mixture that would be too leanto burn in a typical spark-ignited engine. Applicants have determinedthat PCCI combustion does not initiate at, and propagate out from, asingle location. On the contrary, the results show that combustionincludes multiple ignition sites distributed throughout the combustionchamber.

[0128] For efficient, low emission PCCI combustion, it is important tohave combustion occur during an appropriate crank angle range during theengine cycle. If combustion starts too early, cylinder pressures will beexcessively high and efficiency will suffer. If combustion starts toolate, then combustion will be incomplete resulting in poor HC emissions,poor efficiency, high carbon monoxide (CO) emissions, and poorstability. Applicants have determined that the timing of the SOC and thecombustion rate, and therefore combustion duration, in a PCCI engineprimarily depend on the temperature history; the pressure history; fuelautoignition properties, e.g. octane/methane rating or activationenergy, and trapped cylinder charge air composition (oxygen content,EGR, humidity, equivalence ratio etc.). The present invention presents astructured approach to affecting these variables in such a way that thestart of combustion and/or the combustion rate (heat release rate) canbe controlled through various combinations of features discussed morefully hereinbelow.

[0129] The various control features for controlling the start ofcombustion and the combustion rate are controlled/varied to ensureoptimum combustion throughout engine operating conditions so as toachieve low NOx emissions and high efficiency. Application of thesecontrol features will cause combustion to occur within a preferred crankangle range relative to the top dead center position of the enginepiston. Specifically, applicants have recognized that substantially allof the combustion event should occur between 20 crank angle degrees BTDCand 35 crank angle degrees ATDC. Also, combustion would be initiated,preferably between 20 crank angle degrees BTDC and 10 crank angledegrees ATDC, and ideally, approximately between 10 degrees BTDC and 5degrees ATDC. In addition, the duration of the combustion event willtypically correspond to a crank angle in the range of 5-30 crank angledegrees. Preferably, however, one or more of the control features listedbelow will be controlled to prolong the duration of combustion toapproximately 30-40 degrees to achieve desirable peak cylinder pressuresand reduced noise. Thus, optimal control of one or more of the followingfeatures will effectively control the start of combustion and/or therate of combustion such that substantially all of the combustion eventoccurs between 20 crank angle degrees BTDC and 35 crank angle degreesATDC. Of course, there may be conditions under which the start ofcombustion occurs outside the above-stated crank angle range and/or theduration of combustion in the PCCI engine occurs over a broader crankangle range, or may extend beyond the limit described above.

[0130] Applicants have shown that stable, efficient PCCI combustion canbe achieved with most of the heat release occurring after TDC. Forexample, as shown in FIG. 2, the centroid of heat release may bepositioned at 5° ATDC. Applicant have determined that, at light load andlean conditions, as shown in FIG. 3, heat release duration may be in therange of approximately 21.5-25 crank angle degrees.

[0131] As shown in FIGS. 4a, 4 b and 4 c, applicants have determinedthat with an engine running close to its misfire limit, the SOC and endof combustion (EOC) progressively retard and heat release durationlengthens. Gross indicated mean effective pressure (GIMEP) passesthrough a maximum as the SOC retards to after TDC. Meanwhile, the knockintensity and peak cylinder pressure (PCP) decrease substantially closeto the misfire limit, while GIMEP remains acceptable. As shown in FIG.5, the peak heat release rate also decreases and the heat releaseduration increases as the misfire limit is approached. Meanwhile, asshown in FIG. 6, the peak cylinder pressure decreases as the heatrelease rate retards. Clearly, the engine cannot sustain this reactionprocess without providing the appropriate controls discussed herein.Applicants have determined that the best operating point occurs with theSOC occurring a few degrees after TDC. Certainly, improving thePCP-GIMEP tradeoff for PCCI combustion requires a SOC after TDC. As aresult, it is clear that variable, active control is necessary tomaintain the SOC and duration of combustion at the desired location andat the desired length, respectively, to achieve effective, efficientPCCI combustion.

[0132] Variation in the SOC, between sequential combustion events in asingle cylinder engine and between cylinders in a multi-cylinder engine,is due to the sensitivity of PCCI combustion to the pressure andtemperature history leading up to the particular combustion event. Verysmall variations in the compression ratio, the amount of trappedresidual, wall temperatures, etc. have a significant effect on thepressure and temperature history. The present PCCI engine and method ofoperating the engine include control variables/features capable ofcompensating for, and controlling, these variations to achieve optimumPCCI combustion.

[0133] Generally, the control variables, which can be used toeffectively control the commencement of combustion and the combustionrate so as to ensure that substantially all of the combustion processoccurs within the optimal crank angle limit, i.e. 20 degrees BTDCthrough 35 degrees ATDC while minimizing emissions and maximizingefficiency, may be classified in four categories of control: temperaturecontrol; pressure control; control of the mixture's autoignitioncharacteristic; and equivalence ratio control.

[0134] Temperature Control

[0135] The temperature of the in-cylinder air/fuel mixture (in-cylindertemperature) plays an important role in determining the start ofcombustion. The in-cylinder temperature may be varied to control thestart of combustion by varying certain key control features, such ascompression ratio (CR), intake manifold temperature (IMT), exhaust gasrecirculation (EGR), residual mass fraction (RMF), heat transfer andtemperature stratification.

[0136] Applicants have determined that intake manifold temperature (IMT)has a significant effect on propane-fueled PCCI combustion. During twoof Applicants'studies, engine speed, equivalence ratio (Φ) and intakemanifold pressure (IMP) were held constant while IMT was swept throughthe practical operating range. The lowest IMT was limited by unstableoperation and the highest IMT was limited by maximum allowable peakcylinder pressure (PCP). The conditions of the first and second studies,respectively, included engine speed=1200 rpm and 2000 rpm; equivalenceratio =0.30 and 0.24; and IMP=3.3 bar and 4.1 bar. As shown in FIGS. 7aand 7 b, increasing IMT resulted in increased GIMEP and a decreasedcoefficient of variation (CoV) of GIMEP. Also, increasing IMT increasedthe PCP as shown in FIG. 7c, while advancing the SOC and decreasingcombustion duration (FIGS. 7d-7 f). Increasing IMT also increased grossindicated thermal efficiency (FIG. 7g) and the estimated noise (FIG.7k). With respect to emissions, increasing IMT decreased FSHC emissions(FIG. 7h), decreased fuel specific carbon monoxide (FSCO) emissions(FIG. 7i), but had no observable effect on FSNOx (FIG. 7j).

[0137] In summary, Applicants have determined that small changes in IMThave large effects on many aspects of propane-fueled PCCI combustion. Byvarying the intake temperature, the combustion event can be advanced orretarded. Increasing the intake temperature will advance the start ofcombustion; decreasing the intake temperature will retard the start ofcombustion, as shown graphically in FIG. 8. This temperature control maybe accomplished using heat exchangers or burners. For example, a chargeair cooler may be positioned along the intake manifold. A burner orheater in combination with a cooler offers exceptional intaketemperature control. The exhaust products of the burner may be directlymixed with the intake air, the burner could use the intake air directlyfor its air supply, or the heat generated by the burner could be addedto the intake air through a heat exchanger. The heat exchanger may usewaste heat in engine coolant or exhaust gases to heat the intake air.Also, rapid control of IMT can be achieved by using a charge air coolerbypass. A regenerator (similar to that used in a Stirling engine) couldbe used to recover and transfer exhaust heat into the intake air througha heat exchanger thereby controlling the intake temperature. Inaddition, IMT could be varied by injecting fuel into the manifold indifferent phases, e.g. as a liquid or a gas. The change in the heatrequired for vaporization of a liquid fuel would reduce IMT. Of course,different types of fuels would have different effects on IMT.

[0138] Applicants have also determined how residual and intaketemperature, boost and combustion chamber and port wall heat transfer,affect in-cylinder bulk temperature throughout intake and compression,and also the effect on spatial temperature distribution at TDC.Specifically, Applicants compared the intake and compression events foran engine running on an air and propane mixture. Applicants determinedthat the temperature at the SOC is also determined in part by thereheating of the intake charge by existing heat energy. For the purposesof this application, reheat is defined as T(average in-cylinder @ intakevalve closing (IVC)) —T(average intake manifold), that is, thedifference between intake manifold temperature, i.e. temperatureassigned at the inlet to the port and the in-cylinder bulk temperatureat IVC. Applicants determined that reheat starts in the port andcontinues in-cylinder. Moreover, 56% of reheat was due to wall heattransfer and 44% due to mixing and boost for the condition examined.Clearly, heat transfer is very important in determining reheat.

[0139] One study that elucidates the importance of the wall temperatureson the in-cylinder heat transfer is the following. In comparing thefiring cylinder to the misfiring cylinder, it was noted that themisfiring cylinder's reheat was 63% of the firing case (27 vs 43 K).Lower wall temperatures for a misfiring cylinder compared to a firingcylinder are the main reason for its lower in-cylinder temperatures. Thefiring cylinder had a TDC in-cylinder temperature 46 K higher than amisfiring cylinder, compared to a 16 K higher temperature at IVC. Ifcompression were done adiabatically for each case, the temperaturedifference at TDC would have been ˜35 K given the initial 16 Kdifference. Therefore, ˜11 K (46-35 K) temperature loss from IVC to TDCis due to cooler misfiring wall temperatures. Interestingly, althoughwalls heat the in-cylinder gases for the majority of the intake andcompression event, relatively fast rates of heat transfer out of the gasnear TDC compression can result in cooler in-cylinder contents than ifthere were no heat transfer at all. Also, mass flow rate decreased 7.5%due to heat transfer when comparing a normally firing cylinder with wallheat transfer to a firing cylinder with adiabatic walls, primarily dueto the density effect.

[0140] Referring to FIG. 9, with respect to the effect of walltemperatures, i.e. piston temperature, head temperature, and linertemperature, on the SOC, Applicants have determined that as walltemperatures are increased, SOC becomes more advanced. The increasedsurface temperatures cause lower heat transfer to the combustion chambersurfaces thereby advancing combustion. Applicants have shown that withwall temperature varying from 255 to 933 K and all other parameters keptconstant (IMT =342 K, reheat=43 K, φ=0.24), the mixture did not ignitewith a wall temperature below 400 K. From about 400 K to 550 Kcombustion duration increases as a larger percent of the fuel burns.Above 550 K all the fuel burns and the combustion duration decreaseswith increasing temperature. Varying in-cylinder surface temperaturescan be achieved by varying the cooling effect of the engine coolantand/or the lubricating oil on the cylinder/piston assembly. Althoughcylinder wall temperature may be difficult to use as a lever foreffectively controlling SOC, cylinder wall temperatures are one of theparameters considered when controlling SOC, particularly for starting ortransient operation. Applicants have shown that there is a region ofoperating conditions where there are two stable solutions: one withoutcombustion and cool walls, and one with combustion and hot walls. Also,varying the surface to volume ratio in the combustion chamber can changethe heat transfer and, therefore, can be used to control the combustion.

[0141] By comparing a normally firing cylinder with wall heat transferto a firing cylinder with adiabatic walls, wall heat transfer is seen tobe the major contributor to spatial temperature distribution at TDC.Spatial temperature distribution is defined as the manner in which thetemperature varies throughout a region, be it in the port, or in thecylinder at a particular crank angle. By varying the in-cylindertemperature distribution, the start of combustion and/or the overallcombustion rate can be positively affected. One way to vary in-cylindertemperature distribution is to use split intake ports arranged so thatsome of the incoming air/fuel mixture is warmer/colder than the rest ofthe incoming mixture. Another method is to introduce hot spots in thecylinder or to use a glow plug 44 (FIG. 1a). Also, in-cylindertemperature distribution may be controlled by varying the temperature ofthe combustion chamber walls (e.g. the wall temperature of the cylinderliner, piston and/or engine head) by varying, for example, thetemperature of the engine coolant, the temperature of the engine oil orthe rate of cooling of the combustion chamber walls. As shown in FIG.1b, the temperature of the engine coolant may be varied by controllingthe flow through a coolant heat exchanger 46 positioned in the enginecoolant circuit 47 by varying the flow through a bypass circuit 48 usinga bypass valve 50. It was determined that wall heat transfer has similarimpact on spatial temperature distribution for both firing and misfiringcylinders. Similarly, applicants also determined how residualtemperature and wall heat transfer affect in-cylinder temperaturedistribution throughout intake and compression. The determinationincluded three studies of the intake and compression events of an airand propane mixture. These studies revealed that, during most of intakeand compression, hot residual is the main source of spatial temperaturevariation. However, near TDC compression, residual history is of minorimportance compared to heat transfer with the walls in setting uptemperature variations in the combustion chamber. As a result, it isbelieved that to promote a combustion event that uses more of the fuelthat is available, fuel may be introduced in such a way that at SOC,fuel and air exist in proper proportion in regions where the temperaturefield is adequate to sustain combustion. Two areas where the temperaturefield is inadequate to sustain combustion are in the crevices andadjacent cooled surfaces. It is therefore desirable to keep the fuelaway from both the crevices and cooled surfaces.

[0142] Clearly, heat transfer into the in-cylinder mixture increases thetemperature of the in-cylinder mixture thus advancing SOC. Applicantshave shown that a glow plug can be used to effectively control the SOCto a small degree. As shown in FIG. 10, once the glow plug is turnedoff, the SOC and EOC retard slightly. Also, GIMEP decreasessignificantly since less fuel is being burned. The decrease in theamount of fuel being burned also results in a decrease in the heatrelease rate as shown in FIG. 11. Between cycles #1 and #100, the glowplug was turned off and remained off until a time between cycles #300and #400, at which point it was turned back on. Perhaps mostimportantly, when the glow plug is turned off, the start of rapidcombustion is significantly delayed without an increase in duration,which in combination with the decrease in heat release rate, causes thecumulative heat release to decrease. Thus, glow plug 44 (FIG. 1b) couldbe used to positively control combustion to a limited degree.

[0143] In any practical reciprocating engine, heat will be lost from thecombustion chamber during the compression process. The heat loss dependsupon many factors, but primarily upon engine speed and the temperaturedifference between inside and the outside of the cylinder. This heattransfer during the compression process becomes a problem for dieselengines during cold ambient starts as combustion can be difficult toinitiate and sustain in cylinders where the combustion chamber surfacesare cold. Typically, the cylinders located at the ends of each bank ofcylinders run the coldest and are the least likely to fire. It is quitecommon under such conditions for the charge in the end cylinders to failto combust due to excessive heat exchange with the colder cylinderwalls. With diesel engines, however, once all the cylinders warm up,combustion is quite consistent and much less dependent on combustionchamber surface temperatures.

[0144] With PCCI, the combustion process is initiated by obtaining acertain pressure and temperature “history”. Thus, as discussedhereinabove, the PCCI combustion process is strongly dependent upon, andsensitive to, the surface temperatures of the combustion chamber. Thepresent PCCI engine may include an end cylinder compensating means forachieving desired combustion chamber surface temperatures in the endcylinders to ensure better cylinder-to-cylinder temperature controlthereby increasing the likelihood of stable combustion and very low NOxemissions. The end cylinder compensating means may include a system forreducing the effective cooling of specific cylinders, such as reducingpiston cooling nozzle flow; increasing coolant temperature; or reducingcoolant flow rate. Specifically, referring to FIG. 12, the end cylindercompensating means may include an oil flow control system 70 includingoil flow control valves 72 positioned in branch flow passages 74delivering cooling oil to piston cooling nozzles 76 from an oil pump 78.Thus, control valves 72 can be controlled to vary the flow of coolingoil to the piston assemblies to vary the temperature of the piston andthus favorably influence the in-cylinder temperature. Alternatively,flow restrictions could be used instead of valves 72, or the nozzles 76associated with the end cylinders may be designed with a smallereffective flow area than the remaining nozzles to permanently reduce theflow to these piston cooling nozzles. In addition, if more than onenozzle 76 is provided as shown in FIG. 1a, the number of nozzlesoperating could be varied by controlling the respective control valvesassociated with each nozzle.

[0145] Referring to FIG. 13, end cylinder compensating means may includean engine coolant flow control system 80 including a coolant pump 81 andcoolant flow control valves or restrictions 82 positioned in branchpassages 84 leading to the end cylinders 86 of the engine 88. The valves82 are operated to reduce the flow of cold coolant delivered from aradiator 90. Also, control valves 92, positioned in hot coolant returnpassages 94, are used to control the flow of higher temperature coolant,bypassing radiator 90, and delivered directly to the end cylinders.These systems all function to control the flow of coolant to the endcylinders to compensate for the fact that they are cooled more by theambient surroundings so that the total cooling to each end cylinder isequal to each of the other cylinders. These systems can be used toassist in cylinder warm-up to improve engine startability and to provideenhanced control of cylinder combustion and cylinder-to-cylinderbalancing.

[0146] The end cylinder compensating means may, alternatively, oradditionally, include end cylinders having an effective compressionratio nominally greater than the other cylinders to offset the extraheat loss. This compression ratio could be designed into the endcylinders so that the end cylinder compression temperature is equal tothe middle cylinders. This approach is advantageous from a performanceperspective since end cylinder combustion chamber surface temperatureswould be enhanced for both start-up as well as warmed-up operation. Thiscompression ratio difference may alternatively be accomplished throughthe camshaft valve lobe phasing. In this scenario, the end cylinderswould have intake valve closing (IVC) near bottom dead center (BDC) sothat the effective compression ratio (CR) is approximately equal to thegeometric CR. The middle cylinders could then have a retarded IVC whichwould produce a lower nominal effective CR than the end cylinders. Theeffect of varying the compression ratio on PCCI combustion is discussedmore fully hereinbelow.

[0147] One of the biggest challenges with premixed charge, compressionignition (PCCI) engine technology is in the placement of the heatrelease profile. Start of combustion with standard diesel or sparkignition engines is controlled with injection timing or spark timing.With PCCI engines, the start of combustion is dictated by thein-cylinder temperatures and pressures. As SOC timings near TDC (andafter) are approached on the PCCI engine, the sensitivity to smallgeometric and/or operational variations in temperatures, pressures, etc.increase dramatically. As retarded heat release profiles are sought forPCCI engines (to minimize peak cylinder pressures and improveefficiency), the risk of misfire or partial burn increases dramatically.This is due to the fact that the cylinder temperatures decrease aftertop dead center due to the expansion of the charge. If autoignition hasnot yet occurred by TDC, autoignition will not likely occur much aftertop dead center. This problem is further aggravated if one cylinderbegins to misfire. The misfiring cylinder cools down making it even morelikely that the misfiring will continue.

[0148] In a multi-cylinder engine variations inevitably exist betweencylinders with respect to compression ratio, wall temperatures, reheatand residual mass fraction. This variability makes it quite difficult tooperate a PCCI engine with the desired retarded combustion timing whilemaintaining optimum combustion without having individual cylinders(which happen to be running slightly cool) begin to misfire.

[0149] Applicants have determined that manipulating valve events canhave a significant effect on the temperature at TDC and therefore is aneffective tool for controlling the start of combustion as suggested byanalytical results shown in FIG. 14. Specifically, referring to Table I,varying valve events has the following effects: TABLE I effect ofadvancing valve effect of retarding valve modified timing relative totiming relative to event baseline baseline baseline EVC −357° traps hotresidual which exhaust blown back into advances SOC intake whichadvances SOC EVO   135° no effect no effect IVC −167° Miller cycle -lowers at these particular effective CR which conditions, retardingretards SOC slightly improves breath- ing; retarding further reduceseffective CR which retards SOC IVO   341° allow hot exhaust to flowrestricts flow from intake into intake which manifold which has advancesSOC minimal effect on SOC

[0150] As shown in FIG. 15, exhaust valve closing (EVC) plays asignificant role in determining the amount of combustion products thatremain in, or are made available to, the combustion chamber from onecombustion event to the next event, i.e. the residual mass fraction(RMF). The residual exists at a higher temperature than the incomingcharge and therefore heats the charge for the next combustion event.Thus, the timing of exhaust valve closing can be used to adjust thein-cylinder temperature and therefore controlling the SOC. In order to“heat up” a cold cylinder (e.g. one that is beginning to misfire) theresidual mass fraction can be increased in the individual cylinder by anearly exhaust valve closing event. These hot residuals will increase thereheat of the incoming charge and tend to advance the start ofcombustion thereby, for example, restoring a misfiring cylinder. Asshown in FIG. 15, advancing EVC traps hot residual in the cylinder whileretarding EVC allows hot exhaust to be blown back into the cylinder (inthis case, exhaust manifold pressure (EMP)>IMP). The baseline EVC is theoptimum of these two effects: trapping the minimum amount of residualand resulting in the lowest TDC temperature. Similarly, advancing IVOallows some of the hot residual in the cylinder to be blown back in tothe intake, again because EMP>IMP, causing the TDC temperature toincrease. Lowering compression ratio, discussed more fully hereinbelow,by, for example, advancing IVC, will also increase residual in thecylinder, but to a lesser extent. Adjusting the timing of exhaust valveclosing may also be used to effectively compensate for the smallgeometric and operational variations between the cylinders to permit theengine to be “tuned” cylinder-to-cylinder. Any other means foreffectively increasing or decreasing the RMF may be used to advance orretard the SOC, respectively.

[0151] One method of implementing this strategy has been successfullytested on a multi-cylinder PCCI engine. This technique involved theincrease of the exhaust valve lash setting. Opening up the lasheffectively closes the exhaust valve early and advances the start ofcombustion as desired. Applicants have determined that reducing theexhaust valve event by 10 degrees leads to slightly higher surfacetemperatures and 22 degree warmer inlet temperatures. Given the dramaticeffect that 22 degree IMT swings have on combustion (FIGS. 7c-7 f), thismethod would indicate a potential for tuning the multi-cylinder enginewith valve lash adjustments. As shown in FIG. 16, shortening theduration that an exhaust valve is open by increasing the lash doesindeed advance combustion. Ultimately, cylinder-to-cylinder variationscan be controlled passively by any means which can adjust the staticexhaust valve closing. It could also be controlled actively if it iscoupled with some diagnostic measurements. If control exists on allcylinders then this technique could also be used to effect the overallstart of combustion within the engine.

[0152] Another method of controlling in-cylinder temperature bycontrolling the residual mass fraction (RMF) is to compress a pocket ofresidual gas from the previous cycle in a chamber positioned separatefrom the incoming charge. The proportion of trapped residual to freshcharge can be manipulated by the size of such a chamber. The mass of hotexhaust could be as large as (½)(1/CR) and therefore. {fraction (1/30)}of the chamber mass if all the TDC volume is in such a chamber. Thestructure of such a chamber will have to be managed to make at least aportion of the hot gas survive the compression process withoutcompletely mixing with the incoming charge. If the trapped exhaust ismixed very early in the compression process, the high temperaturerequired to initiate the fast reactions will not be reached. The timingof flows into and out of such a chamber may help manage the timing ofthe beginning of rapid energy release in cylinder. Additional sources oflocal heat input may be able to supply such a fast reaction initiation.This might be a heated glow plug or a thermally isolated mass.

[0153] The residual mass fraction is also sensitive to the exhaustmanifold back pressure (EMP). By increasing EMP relative to IMP, theresidual mass fraction can be increased thus increasing the temperatureof the charge which, in turn, advances combustion. Applicants havedetermined that raising EMP does have the expected result of advancingSOC. However, applicant also showed that SOC advanced only by about 4°with a 3 bar increase in EMP for a four cycle engine. Applicants havedetermined that the increase in temperature is nearly linear withincrease in EMP, with all other things being held constant. For a 1 barincrease in EMP, temperature at TDC increased about 10 K. Therefore,considering the practical range of EMP, controlling EMP seems to be arelatively weak lever in controlling SOC on a four cycle engine.Moreover, a very substantial BSFC penalty is paid when using EMP toincrease TDC temperature in a four cycle engine. The BSFC would besignificantly higher than using either exhaust valve closing or variablecompression ratio. Although the effect of increasing EMP is the same asadvancing EVC, i.e. trapping more hot residual mass in the cylinder, theBSFC is much higher because, when EMP is increased, the piston has towork against that pressure through the entire exhaust stroke. If theengine has turbomachinery, then further complications would arise withtrying to use EMP to control SOC. However, using an exhaust restrictionmay still be viable on a two-cycle engine.

[0154] Another important way to control intake temperature is by usinghot exhaust gas recirculation (EGR). As shown in FIG. 1b, a highpressure EGR circuit 54 may be used to direct hot exhaust gas fromupstream of turbine 25 into the intake system 23. EGR circuit 54includes a high pressure EGR control valve 60 for controlling therecirculation of exhaust gas. A low pressure EGR circuit 62 and controlvalve 64 may be used to direct a flow of low pressure EGR fromdownstream of turbine 25 into the intake system 23. Applicants haveshown that EGR is especially effective in increasing the intake manifoldtemperature when introduced upstream of the compressor 24 (assuming theeffect of adding EGR is not cancelled by additional charge air cooling).Exhaust gas recirculation (EGR) has more utility in PCCI engines becausethe exhaust gas of such an engine will contain less particulates andthus the exhaust gas can be recirculated to the ideal upstream location(intake of compressor of turbocharger). The intake of the compressor isthe best location because the pressure differential is almost alwaysfavorable. The fresh intake air and hot EGR mixture will get compressedby the compressor thereby providing heating and mixing. By introducingthe EGR upstream of the compressor and increasing the compressor inlettemperature, the result is a much higher compressor outlet temperaturethan if the EGR is introduced after the compressor. Introducing EGR intothe intake of the compressor is very difficult in normal diesel enginesbecause the particulates in the exhaust gases of the engine “gum up” thecompressor. In a PCCI engine, however, the virtually particulate freeexhaust could be introduced upstream of the compressor withoutsignificant problems. Also, as shown in FIGS. 16, 17 and 18, applicantshave determined that, regardless of the technique used to introduceexhaust products, e.g. EGR, RMF, etc., by adding exhaust products whilemaintaining the temperature of the charge by, for example, injecting acooling diluent, such as air and/or water, the combustion rate can beslowed thus increasing the combustion duration, retarding combustion anddecreasing the amount of heat release.

[0155] Referring to FIG. 19, an improved engine 100 is shown whichbenefits from the PCCI engine and control system of the presentinvention by operating a limited number of a plurality of cylinders in aPCCI mode while operating the remainder of the cylinders in a dieselmode. Specifically, for example, five cylinders 102 in a six cylinderengine may be operated in the diesel mode while one cylinder 104 isoperated in a PCCI mode. This engine also includes an EGR system 106associated only with the PCCI cylinder 104 and separate from an exhaustsystem 108 associated with the diesel cylinders 102. The pressure of thepiston in the PCCI cylinder 104 is used to force the exhaust gas intothe intake system. The EGR system 106 includes an EGR cooler 110utilizing, for example, engine coolant, which cools PCCI exhaust gasbefore recirculating the gas to the upstream side of a compressor 105.Of course, the exhaust gas could be delivered to the intake manifold 112serving only diesel cylinders 102. A well known problem confronted inthe use of EGR in diesel engines is the excessive amounts ofparticulates and NOx present in diesel engine exhaust gas. The improvedengine 100 permits a diesel engine to benefit from EGR whilesubstantially avoiding the drawbacks associated with heavy particulatediesel exhaust thereby providing a less complex and costly system. Forexample, as discussed hereinabove, the PCCI EGR from cylinder 104 couldmore easily be introduced upstream of the compressor without fouling thecompressor. Also, the low NOx emissions of the PCCI EGR reduce theformation nitric acid thereby reducing corrosion in the engine.Applicants have shown that the engine of FIG. 19 lowers the brakespecific NOx emissions while only negligibly increasing the brakespecific fuel consumption.

[0156] Perhaps one of the most effective control features for varyingthe temperature at TDC and therefore the SOC is variable control of thecompression ratio (CR) of a cylinder. By varying the effective or thegeometric compression ratio, both the temperature and the pressurehistories can be controlled. Increasing the compression ratio advancesthe combustion event. Decreasing the compression ratio retards it. Forcertain purposes, the compression ratio may range from 24:1 (to promotecold starting) to 12:1 (to permit control over the start of combustionand limit the peak combustion pressures). The range of compressionratios would depend on, among other factors, the type of fuel used (morespecifically its autoignition properties), for example, natural gas orpropane. Applicants have determined the effect of compression ratio onPCCI combustion. For example, referring to FIG. 20, applicants haveshown that varying the compression ratio is a large lever in changingin-cylinder temperature and therefore SOC. As shown in FIG. 21,applicants have shown that variations in compression ratio significantlyaffects the location of the SOC relative to TDC.

[0157] The compression ratio can be varied by varying the geometriccompression ratio, i.e. using a control mechanism to vary the physicaldimensions/shape of the combustion chamber. The present inventionincludes a compression ratio varying device 38 for varying the geometricor the effective volume of the combustion chamber during engineoperation to achieve a desired SOC. The compression ratio varying devicemay be a mechanical device for causing compression heating of the chargenear TDC by changing the geometric volume of the combustion chamber. Asshown in FIGS. 22a-22 d, the compression ratio varying device mayinclude a movable auxiliary piston or plunger which moves to extend intothe combustion chamber at a crank angle near TDC to decrease thecombustion chamber volume thereby increasing the compression ratio andheating the charge sufficiently to allow ignition to start. The keyfunction of the plunger is to displace some of the charge near TDC.Therefore, the shape and location of the plunger in the combustionchamber will not be critical to its function, except to the extent thatthe plunger affects the crevice volume.

[0158] The size of the plunger will be based on the desired compressionratio control range and may be estimated by the following example:

[0159] Swept volume (displacement) per cylinder=1,000 cc=11.

[0160] TDC clearance volume=100 cc

[0161] Compression ratio=(1000 cc+100 cc)/100.0 cc=11.0

[0162] If the plunger volume=30 cc, then the effective compression ratiowith plunger fully extended=(1000 cc+100 cc)/(100 cc−30 cc)=15.7.

[0163] For a given set of conditions, the modified compression ratioshould be sufficient to allow a large enough increase in temperature andpressure to cause compression ignition for a fuel/air mixture that wouldnot ignite without the plunger. Of course, the engine's compressionratio and the size of the plunger are easily changed during the designstage of the engine. Also, different fuels and intake temperatures couldrequire different plunger sizes and compression ratios.

[0164] As shown in FIG. 22a, the plunger 150 may be positioned in a bore152 in the cylinder head 154 and operated by a cam 156 rotated inpredetermined timed relationship to the movement of the engine piston158. A retraction spring 160 biases the plunger toward cam 156 toincrease the size of combustion chamber 162. This particular arrangementis advantageous in that cam driven plunger 150 can put work back intothe camshaft as the plunger retracts. Also, some of the work thatplunger 150 does on the charge can be extracted by the engine piston, aslong as plunger 150 does not retract until late in the expansion stroke,or after the expansion stroke.

[0165] Alternatively, referring to FIG. 22b, a plunger 170 may behydraulically operated by a pressurized supply of fluid, e.g. fuel,delivered to a chamber 174 by a hydraulic circuit 172 connected to, forexample, a jerk pump or common rail system. FIG. 22c, illustratesanother hydraulically actuated embodiment in which a plunger 180 isassisted by a spring 182, positioned in a chamber 184 formed adjacentone end of plunger 180, to allow energy to be stored in the spring. Inthis system a retaining mechanism, e.g. hydraulic, electromagnetic ormechanical, (not shown) maintains the plunger in the unextendedposition. When the piston is near TDC, a hydraulic fluid supply system186 forces plunger 180 down (at this point the retaining system nolonger holds the plunger). This downward motion is heavily assisted byspring 182. After combustion, plunger 180 moves back up recompressingspring 182 thereby returning energy to the spring. To optimize thisenergy extraction process, the hydraulic chamber 184 bleeds down at arate controlled by a valve 188.

[0166]FIG. 22d illustrates yet another embodiment in which a spring 190,biasing a plunger 192 into the extended position, is strong enough toovercome the gas pressure in the combustion chamber before combustion.Near TDC, a bleed down valve 194, connecting a chamber 196 is opened andthe spring 190 pushes plunger 192 into the extended position in thecombustion chamber 162 causing the charge to ignite and the pressure inthe combustion chamber 162 to increase. As a result, plunger 192 ispushed back up against spring 190. If needed, a high pressure supply 200supplies hydraulic fluid to chamber 196 to ensure plunger 192 moves backup into the retracted position. If the gas pressure is sufficient tomove the plunger back up into the retracted position, a low pressurehydraulic fill supply 202, including a one-way valve 204, may be used tofill the chamber 196 below plunger 192.

[0167] The compression ratio may also be varied by providing an opposedpiston engine design having variable phase shifting to permit thecompression ratio to be varied during operation by changing the phase ofrotation between two crankshafts. The opposed piston engine may be ofthe type disclosed in U.S. Pat. No. 4,010,611 or of the interconnectedcylinder type with variable phasing as disclosed in U.S. Pat. No.4,955,328, the entire contents of both of these references being herebyincorporated by reference. Alternatively, referring to FIG. 23, thecompression ratio could be varied using a phase shifting mechanism 210including a conventional differential assembly 211 connected between aninput shaft portion 212 of one of the crankshafts 214, 216 associatedwith respective pistons 218, 220 and an output shaft portion 222 of thesame crankshaft 214 to permit the portions of the crankshaft to berotatively shifted relative to one another. Crankshafts 214 and 216 areconnected via a conventional gear assembly 223 for transferring power toa driven shaft 225. As shown in FIG. 24, the differential 211 includes aring gear 224 mounted on one end of input shaft portion 212, an arm 226extending from ring gear 224 and a gear assembly 227 mounted on theopposing ends of shafts portions 212, 222. A rotator mechanism 228,including a pinion gear 230, is operatively connected to ring gear 224to rotate the ring gear when a change in the phasing between thecrankshafts is desired. As long as ring gear 224 remains stationary,shafts portions 212, 222 remain in phase. When ring gear 224 is rotatedby rotating pinion gear 230, arm 226 rotates causing a change in thephasing between shaft portions 212, 222. The rotator mechanism 228would, therefore, be used to adjust the relative phasing of the inputshaft to the output shaft, thereby adjusting the phasing of the twocrankshafts and the compression ratio. In addition, two crankshafts percylinder could be used to eliminate the inherent side thrust imparted bythe crankarm in the single crankshaft design. The effect of the maximumpossible compression ratio on the sensitivity to CR on phasing should benoted. It might be advantageous to have a geometry where the pistonsinterfere with each other at “zero” phasing. Of course, this set upwould operate with non-zero phasing all the time.

[0168] Applicants have determined how the change in phasing of anopposed piston engine changes the compression ratio. This effortincludes three studies as shown in FIG. 25. In the first, when the twopistons were in phase, i.e. both pistons reach TDC at the same time, thecompression ratio was 25:1. In the second, when the pistons were inphase they would come together and just touch at TDC. With a flat toppiston there would be no volume between the pistons and, assuming nocrevice volume, the compression ratio would become infinite. The thirdcase assumes negative interference so that the pistons would come incontact while out of phase to some degree. For this case, the overlapwas about 10% of the stroke causing the pistons to contact at 46° out ofphase. Of course, engine geometry (bore, stroke, connecting rod length)will also effect CR versus phasing; these values were kept constant inthis study.

[0169] These results indicate that the compression ratio could be variedover a very large range using an opposed piston arrangement withvariable phasing. Also, the slope of the change in compression ratiowith phasing depends on the amount of clearance or negative clearancebetween the pistons at TDC with 0° phasing. Thus, in a practicalapplication, it would be desirable to strike a balance between the rangeof phasing needed to cover the desired range of compression ratio andthe precision with which the phasing needs to be controlled, i.e. theslope of the curve in FIG. 25 should be optimized. Thus, ideally, theslope of the curve would be steep enough that the desired range ofcompression ratio could be achieved within a limited amount of phasing,and not so steep that the phasing needs to be too precise.

[0170] Referring to FIG. 26, it is very clear that as the pistons becomemore and more out of phase that the compression ratio decreases. It isalso clear that there is very little change in shape of the cylindervolume versus crank angle curve for phasing angles less than about 120°.As a result, the variation in phasing can be used to control compressionratio over a large range without any affect in the cylinder volumeversus crank angle. An opposed piston system with variable phasingclearly provides the desired flexibility to achieve a broad range ofcompression ratio values.

[0171] The effective compression ratio may be varied with variable valvetiming. Specifically, as shown in Table I, advancing of the intake valveclosing lowers the effective CR while significant retarding of the IVCalso reduces effective CR. However, changing valve events can have avery large effect on the breathing of an engine, and thus the air/fuelratio, in comparison to varying the geometric compression ratio(assuming that the fuel flow rate is held constant). The steepest changein airflow with TDC temperature is when IVC is changed. As WVC becomesearlier, the TDC temperature is lowered, but airflow is severelyrestricted possibly undesirably changing the equivalence ratio. In thiscase, an increase in boost accompanying earlier IVC could be used tomaintain a constant air flow rate. Similarly with EVC, as EVC is changedthe amount of residual trapped in cylinder changes, and thereforebreathing is affected. The slope of the IVC line is roughly twice thatof EVC and IVO while varying the geometric compression ratio does nothave an effect on airflow. In terms of changing TDC temperature withouteffecting airflow, variable geometric compression ratio appears to bethe most effective of the control features.

[0172] Referring to FIG. 28, changing any of the valve events orcompression ratio has a definite effect on BSFC. In order to get thebest BSFC, increasing compression ratio would be a better choice thanchanging exhaust valve closing when a higher temperature is needed. Avery large BSFC penalty would be paid if EVC is advanced to increase thetemperature at TDC. If a lower temperature is needed, advancing IVC isthe best method whereas varying the geometric compression ratio couldalso be an option since it results in only somewhat higher BSFC.

[0173] Applicants have also determined, as expected, that changing theeffective compression ratio has a large effect on peak cylinderpressure, as shown in FIG. 29. IVC has an almost identical curve as VCR,confirming the fact that changing IVC really changes the effectivecompression ratio. Because the heat release starts 5° ATDC in this case,the cylinder pressure trace appears to be “double-humped”: the firstpeak at TDC is due to compression; the second peak after TDC is due tocombustion. The appearance of two slopes of VCR and IVC lines is due tothe absolute peak cylinder pressure occurring either on the combustionhump (CR<18) or the compression hump (CR>18). In order to increase thetemperature at TDC from the baseline without undesirably affecting peakcylinder pressure, changing EVC or IVO would be the best strategy.However, this strategy may result in an undesirable increase in BSFC(FIG. 28) and may also change the engine breathing (FIG. 27).

[0174] Applicants have also determined that very high compression ratiosare needed for combustion at low intake temperatures. For example, ithas been found that at intake temperatures of

[0175] and 40° F., no combustion occurs when the correspondingcompression ratios are below 35, 33, and 30, respectively. At warmed upconditions, the desired compression ratio is approximately 15, whichmeans that a change of approximately 20 compression ratios would beneeded to cover these conditions. Due to the very high compressionratios required under these conditions, peak cylinder pressures are alsohigh and in some cases greater than 200 bar. As a result, intake airheaters and/or some other method of starting in cold conditions may bemore practical than using variable compression ratio alone. Also,maintaining a lower compression ratio will allow a higher GIMEP to beachieved before hitting the peak cylinder pressure limit.

[0176] Another method of controlling the temperature is to introducewater into the intake manifold or directly into the cylinder. Applicantshave shown that when the nitrogen in the intake air is completelyreplaced with water, the water will likely result in a lower flametemperature (205 K lower) due to dissociation. Also, inapplicants'study, the ignition delay increased slightly (by 0.04 msec)and the peak reaction rate dropped by about 50%. Also, when water wasadded into the intake manifold, e.g. water fumigation, the chemicaleffect, although small, is to slightly retard the SOC. However, liquidwater injection into the intake manifold effectively cools the intakemanifold due to the vaporization of the liquid to steam. As a result,IMT and TDC temperatures are significantly decreased as shown in FIG.30. The impact of water injection on temperature at TDC is mostly due tothe decrease in IMT, not due to the change in the ratio of specificheats. The effect on IMT should be viewed as an upper limit.

[0177] It should be noted that applicants have shown that PCCI can bemaintained without adverse thermal effects on the piston 14 (FIG. 1a).Even though PCCI combustion can create knock intensity levels 10-20times higher than the safe level experienced in spark-ignited engines,both aluminum and steel pistons do not reach excessive temperaturelevels. In applicants'preferred embodiment, the temperatures resultingfrom autoignition in PCCI combustion are much lower than thetemperatures experienced in spark-ignited engines since, inapplicants'preferred embodiment, PCCI combustion operates under suchlean conditions.

[0178] Pressure Control

[0179] The SOC may also be controlled by controlling the pressure in thecombustion chamber. One way of controlling in-cylinder pressure is touse a compression ratio varying device to vary the pressure in thecombustion chamber. Although varying the compression ratio ultimatelyvaries both the pressure and temperature of the charge, the pressure isdirectly changed. An increase in the compression ratio will tend toincrease the pressure at TDC, and a decrease in compression ratio willdecrease pressure at TDC. Applicants have shown that increasing thein-cylinder pressure advances the start of combustion and decreasing thein-cylinder pressure retards the SOC. Any of the compression ratiovarying devices discussed hereinabove with respect to temperaturecontrol may be used.

[0180] A second way of controlling the in-cylinder pressure is to varyto the intake manifold, or boost, pressure (IMP). The timing of the SOChas been shown to be a function of pressure. Applicants have determinedthe effects of varying IMP on combustion and engine operation. Theengine conditions for one engine study were 1200 RPM, 355.7K<IMT<357.4K,0.256<Φ<0.263. IMP was varied. Maintaining these conditions whileincreasing IMP required increasing air flow and fuel flow. FIGS. 31a and31 b show that the duration of heat release decreases as IMP increasesboth in the crank angle domain and the time domain. FIG. 31d shows thatSOC occurs earlier as IMP increases. FIG. 31c, showing results fromanother study, clearly indicates that increasing the boost pressuresignificantly advances the heat release event. FIG. 31e shows that FSHCemissions decrease as IMP increases, indicating more completecombustion. FIG. 31f shows that GIMEP increases as IMP increases, mostlydue to the increase in complete combustion, and, to a lesser extent,more fuel. FIG. 31g shows that gross indicated thermal efficiencyincreases as IMP increases, partly due to more complete combustion. FIG.31h shows that FSCO emissions decrease as IMP increases, apparently dueto more complete combustion. FIG. 31i shows that FSNOx emissions are notsignificantly affected by IMP. FIG. 31j shows that coefficient ofvariation (COV) of GIMEP decreases as IMP increases. FIG. 31k shows thatPCP increases as IMP increases. FIG. 311 shows that estimated noiseincreases as IMP increases. FIG. 31m shows that as IMP increases,smaller gains in GIMEP cause larger rises in PCP. This effect is due tothe earlier SOC that occurs as IMP increases.

[0181] One study varied the pressure at BDC of the compression stroke.The study was performed using a compression ratio of 14.5:1, an enginespeed of 1200 rpm, a BDC compression temperature of 389 K, anequivalence ratio of 0.3285, and no heat transfer. The fuel used waspropane and the pressure at BDC was varied while all other parameterswere held constant. This study clearly revealed that as pressure at BDCincreases, he SOC becomes earlier. In addition, for BDC pressures lessthan 1.75 bar, less than 10% of the fuel energy was released, while forBDC pressures greater than P=1.75 bar, virtually all of the fuel energywas released. This indicates that the combustion is highly sensitive tochanges in pressure. At very low pressures, very little of the fuelbums, leading to high FSHC emissions. Since none of the fuel is burningat these low pressures, no carbon monoxide is produced. As the pressureincreases (while maintaining IMT constant), a higher percentage of thefuel is burned, which leads to decreased production of carbon monoxideand lower FSHC. Above a certain critical pressure, all of the fuel bumscompletely, leading to extremely low FSHC and FSCO emissions. Also, avery small change in BDC pressure leads to a very large change in peakcycle temperature (PCT). The results of the simulation indicate that atlow peak cycle pressures (PCP), the fuel does not bum. Hence, thepressure peaks at the isentropic compression. As pressure is increased,a higher percentage of the fuel energy is released, causing the cylinderpressure to rise above the isentropic compression pressure. As pressureincreases further, all of the fuel energy is being released and furtherincreases in pressure raise the PCP due to isentropic effects.

[0182] Clearly, varying IMP can be an effective way of controlling theSOC and the duration of combustion. Increasing the IMP tends to advanceSOC while decreasing the duration of heat release. Likewise, decreasingthe IMP tends to retard SOC while increasing the duration of heatrelease. In a typical application, for a constant torque condition, thefuel flow rate would remain virtually constant, and the boost pressurewould be increased to advance the start of combustion or decrease theboost to retard the start of combustion. For example, an air compressor,a turbocharger, a supercharger such as driven by an engine powertake-off, or an electrically powered compressor, could be used. For agiven power level, and, therefore, for a given fuel flow rate, theretypically exists a preferred intake pressure and temperature. At verylow loads, it may be desirable to control the intake manifold pressurewith a throttle 53 (FIG. 1a) in the same way that the intake pressure iscontrolled on a current production spark ignited engine. Throttle 53would also be used when operating a multi-mode PCCI engine in a sparkignited mode as described hereinbelow. Of course, a throttle couldalternatively be located at other locations in the intake system, suchas in the intake manifold.

[0183] Air/Fuel Mixture Autoignition Properties

[0184] Another strategy for controlling the start and duration ofcombustion is to vary the air/fuel mixture autoignition properties. Theautoignition properties of the air/fuel mixture may be controlled byinjecting a gas, e.g. air, oxygen, nitrogen, ozone, carbon dioxide,exhaust gas, etc., into the air or air/fuel mixture either in the intakesystem, e.g. preferably in the port using, for example, injector 42, orin the cylinder directly using, for example, injector 40, therebyproviding control over the start of combustion and the combustion rate.

[0185] Applicants have examined the effect of adding reactive species tothe air/fuel mixture on the combustion process. One study was performedusing an equivalence ratio of 0.3, a temperature at BDC of 389 K,pressure at BDC of 3 bar, and propane as the fuel. The compression ratiowas 14.5, and the engine speed was 1800 RPM. The engine geometry usedwas for a Cummins C series engine. The nitrogen, oxygen, and fuel molefractions were held constant at 0.771, 0.216, and 0.0123, respectively,for all cases. The mole fraction for the reactive species added was0.000411 for all cases. The reactive species examined were H₂, H₂O₂, OH,CO, O, HO₂, H, and O₃. FIG. 32 shows the temperature versus crank angle.Although CO and H₂ advanced the SOC by less than 0.5 crank angledegrees, all other species significantly advanced the SOC, with O₃(ozone) causing the largest change in the SOC. Therefore, smallconcentrations of most common radicals will cause significant changes inthe SOC.

[0186] Thus, applicants have determined that the addition of very smallquantities of ozone advances the SOC by significant amounts. Applicantshave also shown that virtually all of the ozone will be consumed by thecombustion process and that the change in the SOC will diminish as theamount of ozone added increases. Specifically, FIG. 33 illustrates theeffects of additional ozone on advancing the SOC. The increase intemperature indicates the start of the combustion event.

[0187] Given the significant effect additional ozone has on the SOC,ozone can be used in several ways to advantageously control thecombustion in a PCCI engine. First, by adding different amounts of O₃ tothe intake ports, one, several, or all cylinders could have their SOCadjusted. Second, adding O₃ to the intake could be used as a coldstarting aid for PCCI and diesel engines. Third, adding O₃ to theexhaust of an engine would allow a catalyst to light earlier thuspossibly significantly reducing cold start emissions oncatalyst-equipped spark ignited engines, diesel engines and PCCIengines. O₃ could be produced “on board” through a simpleelectrochemical reaction. Ozone generators are commercially available.Also, the ignition delay of a diesel engine could be reduced by addingO₃ to the intake. This would reduce the premixed bum fraction whichwould then lower NOx emissions and reduced noise.

[0188] Applicants have shown that increasing the oxygen concentrationadvances the SOC. However, applicants have determined that oxygenenrichment from 20.7 percent to 21.65 percent will advance the SOC byless than one crank angle degree, and oxygen enrichment from 20.7percent to 23.7 percent will advance the SOC by less than 1.5 crankangle degrees. Therefore, combustion may be controlled to a limiteddegree by modifying the oxygen concentration of the intake air. This maybe done by adding oxygen (or an oxygen rich gas mixture) to the intakeor by selectively removing nitrogen from the intake air (using amembrane for example). Applicants have also shown that increasing thepercent of nitrogen in the intake charge from 78.6 percent to 80.6percent resulted in the retardation of the SOC by less than 2 crankangle degrees at 1800 rpm. It was also noted the same percentageincrease of N₂ in the fresh charge lowers the FSNOx from 0.144 to 0.048grams of NOx per Kg of fuel.

[0189] Another method of varying the effect of oxygen on the combustionprocess is to dilute the mixture with EGR. In one study, an engine EGRsystem was plumbed from the exhaust manifold to the compressor inlet.Because the EGR is mixed in upstream of the aftercooler, and in thepresent study, the aftercooler exit temperature was controlled and heldfixed, the EGR should not have significantly effected the temperature atSOC. During this study, fuel rate and intake manifold temperature wereheld constant. As the EGR rate was increased, exhaust manifold pressuredecreased, which in turn decreased air flow on this turbocharged engine.The fuel rate was held constant, so the fresh equivalence ratioincreased. In spite of the increased equivalence ratio, SOC retarded asthe EGR rate increased, most likely due to the diluent effect of theEGR. As expected, SOC retarded as the EGR rate increased. However, asEGR rate increased, CO and HC emissions also increased. Also, as EGRrate increased, the spread in SOC between cylinders increased. In asimilar study, the SOC was held constant by adjusting IMT. As the EGRrate was increased, exhaust manifold pressure decreased, which in turndecreased air flow. The fuel rate was held constant thus causing theequivalence ratio to increase. In addition, as the EGR rate increasedfrom about 7 to 13% EGR, there was a sharp rise in the cylinder tocylinder variation in SOC. Ultimately, a higher IMT was required tomaintain constant SOC as the EGR rate increased, in spite of an increasein equivalence ratio. This requirement was due to the diluent effect ofincreased EGR on the intake air.

[0190] Another technique for modifying the autoignition properties ofthe air/fuel mixture to control SOC and the duration of combustion is tovary the octane, methane or cetane number of the charge by, for example,by providing two or more fuels have different octane, methane or cetanenumbers. The fuel supply can be either selectively switched between thefuels or the fuels can be mixed. This technique makes it possible toretard or advance the combustion event. For example, a fuel which tendsto autoignite more readily (lower octane or methane number, or highercetane number) could be controllably mixed with a fuel that tends toautoignite less readily (or a fuel that ignites at a high temperatureand a fuel that ignites at a low temperature could be used) to enabledirect control over the timing of ignition and rate of combustion bychanging the ratio of the fuels that are present in the combustionchamber during the combustion event. As shown in FIG. 34, propane,octane and heptane have significantly different effects on the SOC. Thesame effect may be achieved by using a fuel additive, such as acontrolled amount of propane, ethane, or other hydrocarbons, such asengine lubricating oil, that change the autoignition properties of thefuel to advance or retard the start of combustion. Of course, any methodthat changes the fuel's octane/methane number or the activation energyof the fuel can be used to advance/retard combustion. Applicants havedetermined that there is a significant sensitivity of start ofcombustion to octane number. This effect was independent of intakemanifold temperature. Moreover, in one study, the start of combustionwas retarded approximately 7° for an increase in octane number from 80to 100.

[0191] Achieving dynamic control over individual cylinder combustion ina multi-cylinder PCCI engine will be critical to achieving improvedcombustion. Since many of the gases/fluids discussed hereinabove, e.g.fuel, ozone, oil, water, etc. have now been shown to significantlyaffect the SOC and/or rate of combustion, these additives can be used toadvantageously balance combustion between the cylinders in amulti-cylinder engine running on PCCI principles. For example, byinjecting a liquid or gas diluent, such as a less reactive fuel, water,uncooled or cooled exhaust products, air and/or nitrogen either into theintake air or directly into the charge in the cylinder, the SOC can beretarded. Also, by injecting, for example, a more reactive fuel, ozone,oil and/or oxygen into the charge the SOC can be advanced. FIG. 1billustrates one system for balancing combustion between cylinders of amulti-cylinder engine. The system uses port injection of fuel on theengine with two types of supplies per cylinder —supply 32 to injectliquid fuel and supply 34 to inject gaseous fuel. Although the supplies32 and 34 are shown feeding into a single passage for delivery to theintake port, the supplies may include separate delivery passagesconnected to the intake port at different locations. Liquid fuel willdecrease intake charge temperature by the heat of vaporization of theliquid fuel. The temperature at TDC compression, and therefore SOC, canbe controlled by varying the amount of liquid versus gaseous fuel. Also,it should not matter if the liquid vaporizes in the port or duringcompression. The gaseous and liquid fuel can be the same fuel indifferent states, e.g. propane, or different fuels, e.g. gaseous naturalgas and liquid gasoline, such as indolene. It is important that the portinjection system have good separation between cylinders and sequential(timed to intake event) injection is likely to be required. Duringoperation, a cylinder that is “going out” would be given more gaseousfuel and a cylinder that is “too hot” would be given more liquid. Thismethod can be used to achieve about a 20 degree temperature difference.One of the supplies could be lubricating oil or ozone while the othersupply could be a fuel having a high resistance to ignition, e.g. a highoctane number, to permit the SOC to be effectively controlled by varyingthe amount of oil or ozone added to the mixture. Also, by using theengine's lubricating oil supply, or using ozone created by the engineduring operation, an additional supply of fuel/additive can be avoided.

[0192] Equivalence Ratio

[0193] Another control variable that applicants have shown can beeffectively used to control the SOC and combustion duration or heatrelease rate is the equivalence ratio (p of the fuel/air mixture.Equivalence ratio is equal to fuel/air ratio divided by stoichiometricfuel/air ratio (if<1, fuel deficient; if>1, fuel excess). Combustionneeds to be slowed down in a PCCI engine because fast combustion leadsto high noise, lowered efficiency and high peak cylinder pressure. Ifdifferent temperatures and/or equivalence ratios can be achievedthroughout the charge of air/fuel at or near point of ignition, theresulting rate of combustion will possibly be slowed down thusadvantageously lengthening the duration of combustion. The equivalenceratio could be increased by increasing the fuel flow to the cylinderwithout a corresponding increase in intake air flow, or by decreasingthe intake air flow. The equivalence ratio could be lowered bydecreasing the fuel flow to the cylinder without a correspondingdecrease in air flow, or increasing the air flow rate. Variations in thequantity of fuel delivered to a cylinder is varied by controlling theoperation of fuel control valves 39, 41, and/or fuel injectors 35, 36 ina known manner. The variations in the air flow rate could be achievedby, for example, variably controlling compressor 24 to vary boostpressure.

[0194] To test the lower limit for equivalence ratio, applicantsconducted engine studies to determine whether acceptable PCCI combustioncould be obtained with an extremely lean mixture. The results indicatethat very stable combustion can be achieved at an extremely leanequivalence ratio of 0.05 while obtaining a heat release duration ofapproximately 30 degrees. Also, as shown in FIGS. 35 and 36, the resultsindicated that the start of combustion advances and the apparent heatrelease duration decreases as equivalence ratio increases, i.e. air/fuelmixture becomes richer. Applicants have clearly shown, as indicated inFIG. 37, where cylinder temperature increases indicate the heat releaseevent. Moreover, referring to FIG. 38, the apparent heat releaseduration becomes longer as equivalence ratio decreases, i.e. air/fuelmixture becomes leaner. Also, applicant have shown that for a fourstroke engine that both peak cylinder pressure and GIMEP increase asequivalence ratio becomes richer. With respect to a two-stroke engine,applicant have determined that as equivalence ratio increases, GIMEPincreases.

[0195] Studies were also conducted to investigate whether theequivalence ratio affects the amount of fuel burned in PCCI combustion.The results indicated that as equivalence ratio becomes richer, thepercentage of fuel energy showing up as apparent heat released increasesat first and then levels off near 80%. This number can never reach 100%because of heat transfer. With respect to emissions, as equivalenceratio becomes richer, fuel specific hydrocarbon emissions decrease. Inaddition, as equivalence ratio became richer, average noise levelsincreased, and GIMEP increased. As equivalence ratio becomes richer, theaverage knock intensity increases. As equivalence ratio became richer,the cycle-to-cycle combustion variation, as measured by the coefficientof variation (COV) of GIMEP, generally decreased. In fact, the COV's ofGIMEP, for the conditions of the study, stayed below the combustionstability limit (in this case defined as 5%), where a COV above thelimit indicates unacceptable stability.

[0196] Studies were performed to determine the effect that variations inequivalence ratio have on thermal efficiency in PCCI combustion. Anequivalence ratio study was performed while matching the followingparameters: speed, IMT, IMP, engine oil temperature, and engine watertemperature. Equivalence ratio was increased by holding air flowconstant and increasing fuel flow to the engine. As fuel flow increasedand equivalence ratio became richer, gross indicated thermal efficiencyincreased at first and finally leveled off. Engine work output increasedwith respect to increased fuel flow as more fuel was burned. At theleaner equivalence ratios, a significant amount of fuel is leftunburned. At the richer equivalence ratios, the percentage of fuel thatis being burned levels off as noted hereinabove, and the gross indicatedthermal efficiency levels off because the increase in engine output isbeing offset by the additional fuel input.

[0197] In addition, an engine study was conducted with the engine cyclerunning from bottom dead center of the compression stroke to BDC of theexpansion stroke. The study was conducted using a compression ratio of14.5:1, an engine speed of 1200 RPM, a BDC compression temperature of389 K, pressure at BDC of 4.31 bar, and no heat transfer. The fuel usedwas propane. The equivalence ratio was varied while all other parameterswere held constant. It was discovered that the percent of energyreleased slowly tapered off as the equivalence ratio drops below 0.15.This data indicates that for a given temperature and pressure, there isa lower limit to the equivalence ratio of a mixture that will burncompletely. Also, it was shown that FSCO emissions are very high atequivalence ratios below 0.15. This data indicates that only a smallamount of the fuel burns to completion at these low equivalence ratiosfor this temperature and pressure. In addition, the FSHC decreaseslightly as the equivalence ratio is varied from 0.05 to 0.4. Thus, mostof the fuel reacts regardless of the equivalence ratio. It was alsoshown that SOC occurs earlier as equivalence ratio increases. The studyshowed that peak cylinder temperature gradually increases as equivalenceratio is increased showing the increased amount of energy available tobe released. Peak cylinder pressure (PCP) gradually increases asequivalence ratio is increased showing the increased amount of energyavailable to be released. At equivalence ratios greater than or equal to0.18, virtually all of the available fuel energy is released, leading toa nearly linear increase in PCP as equivalence ratio increases.

[0198] Applicants have determined that it may be possible, although notnecessarily desirable, to maintain PCCI combustion at very richequivalence ratios, e.g. 0.5, if IMP and IMT are sufficiently low toprevent the peak cylinder pressure limit from being exceeded. It will bedifficult to start an engine at the low boost and IMT levels needed formaintaining low cylinder pressures at such rich equivalence ratios. Thevery advanced heat release, loud knock, and combustion roughness makerunning at this condition undesirable. A lower CR for retarding SOC mayimprove these aspects.

[0199] Also, by varying the level of charge stratification, thetemperature and equivalence ratio distribution can be altered to permitcontrol of the combustion rate and/or the start of combustion. Anauxiliary combustion chamber concept may be a mechanism for achievingthe desired stratification, thereby enabling better control over thestart of combustion. For example, conventional auxiliary combustionchamber designs typically used on small engines having indirectinjection (IDI), and large spark ignited engines using natural gas fuel,could be used.

[0200] In order to operate under the desired lean conditions for optimalPCCI combustion, substantial air flow must be provided to the intakemanifold. A turbocharger could provide the needed air flow for amulti-cylinder PCCI engine. Applicants'original target was to reach anequivalence ratio of 0.40 or leaner. Referring to FIG. 39, applicantshave shown that operating at leaner than an equivalence ratio of 0.29would violate the compressor pressure ratio limit of the availableturbocharger. Applicant determined that turbine pressure ratios are veryhigh at lean equivalence ratios. As a result, the exhaust manifoldpressure is very high which causes a large BSFC penalty. Because of therelatively cool exhaust temperatures produced by PCCI combustion, verysmall turbine cases are needed which result in high exhaust manifoldpressures.

[0201] Applicants have determined that it would be desirable to operateunder slightly leaner conditions than the original target. At anequivalence ratio less than 0.4, a smaller turbine casing was used todecrease the compressor pressure ratio and exhaust manifold pressureratio, but a high BSFC penalty is paid, as shown in FIG. 40. FIGS. 41and 42 illustrate the higher PMEP losses with the smaller turbine casingand the higher BSFC. Also, with the smaller turbine casing, the rotorspeed is much higher and, in fact, near the limit on rotor speed as seenin FIG. 43 (rotor speed limit 120-125 k rpm range). Applicantsdiscovered that there is a lower limit on the size of the turbine casingused due to the losses incurred with the high back pressure and withreaching the rotor speed limit.

[0202] In order to avoid this problem with the high back pressure androtor speed limiting airflow, one possible solution is to use amechanically driven supercharger in conjunction with a turbocharger. Thesupercharger would be upstream of the compressor so that the turbinebears less of a burden for producing boost. Some BSFC penalty would beincurred for the shaft work absorbed by the supercharger; however, theBSFC penalty is less than the very high penalty incurred with the verysmall turbine. Because the supercharger is driven mechanically from theshaft, there should be no trouble getting the desired air flow. Theturbine then can be sized somewhat larger, and should not approach thespeed limit and should not have extremely high back pressure.

[0203] Applicants have also determined the effect of engine speed onSOC. The time of autoignition depends on the temperature and pressurehistories. By changing the engine speed, these histories are changed. Itis possible to advance the combustion event by reducing the enginespeed, and to retard the combustion event by increasing the enginespeed. Specifically, a 75% increase in engine speed, from 1000 to 1750resulted in a 1.5% increase in the start of combustion pressure and a2.8% increase in the start of combustion temperature. In addition, a 75%increase in engine speed decreased the heat release rate duration by0.81 ms (only a 23% decrease) which corresponds to an increase in heatrelease duration of 1.7 crank angle degrees (only an 8% increase). Giventhis minimal impact of engine speed on the SOC and heat release, and theinability to effectively vary engine speed in many practical engineapplications, engine speed is not viewed as an effective combustioncontrol variable. However, one example where engine speed could be usedto provide some control over combustion is in an application where theengine drives a generator or alternator.

[0204] As discussed hereinabove, the foregoing control variables areused to control the SOC and the duration of combustion to achieveoptimum PCCI combustion. One key consequence of efficient, optimumcombustion is reduced emissions. Applicants have shown that a PCCIengine can achieve NOx emission levels that are well below any other NOxemission levels ever demonstrated by applicants using diesel and naturalgas engines, and well below future emissions standards as shown in FIG.44. The use of propane as the fuel resulted in the lowest NOx emissionsrelative to diesel fuel and gasoline.

[0205] Applicants have also determined the effect of the controlvariables and other factors on emissions of a PCCI engine. Engine speedhas little effect on the quantity of NOx emissions. Although a 75%increase in engine speed approximately tripled the FSNOx, the levels ofNOx emissions produced were still extremely low. Also, as equivalenceratio becomes richer, fuel specific NOx generally increases, but stillremains at extremely low levels. Referring to FIG. 45, applicants havedetermined that engine speed appears to affect FSCO and FSHC emissionsmore significantly. As shown, below a certain critical speed, virtuallyall of the fuel burns, FSHC are low and FSCO is low. Just above thecritical speed, the fuel partially burns, resulting in higher FSCOemissions. As engine speed continues to increase, the percentage of thefuel that burns continues to drop, resulting in lower FSCO emissions.These emissions also vary as the temperature at BDC varies. Referring toFIG. 46, at very low temperatures, very little of the fuel burns,leading to high FSHC emissions. Since none of the fuel is burning atthese low temperatures, no carbon monoxide is produced. As thetemperature increases, a higher percentage of the fuel is burned, whichleads to increased production of carbon monoxide and lower FSHC.Finally, above a certain critical temperature, all of the fuel bumscompletely, leading to extremely low FSHC and FSCO emissions. In fact,as shown in FIG. 47, applicants have shown that all data points with endof combustion flame temperatures above 1600 K had acceptable COemissions. It has been shown that both high temperature and the hydroxylradical (OH) are critical for the desired oxidation of CO. Importantly,as equivalence ratio becomes richer, fuel specific CO decreases, whilethe concentration of CO₂ in the exhaust increases. In one study, allpoints taken with an equivalence ratio<0.2 had CO emissions above theEPA CO limit.

[0206] As equivalence ratio becomes richer, fuel specific HC decreases.Clearly, unburned hydrocarbons (UHC) are a key concern for PCCI enginessince reducing unburned hydrocarbons is essential to the commercialfeasibility of a PCCI engine. Applicants have determined that UHC, andCO, is formed in small crevices positioned in the components forming thecombustion chamber, i.e above the top ring of the piston between thepiston and the liner; between the cylinder head and the cylinder liner;and around the components mounted in the cylinder head. The crevicesprevent the volume of mixture in the crevice from reaching asufficiently high temperature necessary for burning of the HC andoxidation of the CO. For example, applicants have shown that similarpistons with different crevice volumes have different UHC levels. Thepresent PCCI engine may include one of several designs to minimize UHC.The present crevice minimizing designs result in a low crevice volume;keep the fuel away from any existing crevices; or cause the mixture inthe crevice volume to burn appropriately. The designs shown in FIGS. 48aand 48 b are most easily implemented in a ported two stroke cycleengine. Referring to FIG. 48a, in one embodiment, the engine has asingle piece head and liner combination 300, although a two-piece systemcould be used. Just above the top ring 302 (at TDC), the bore 304increases to eliminate the crevice around the top land 306 of piston308. There are no crevices in the cylinder head, as it is a single piecewithout valves, gaskets, etc.

[0207] Referring to FIG. 48b, a second embodiment of the creviceminimizing design may similarly include a one-piece head and liner 310.However, in this embodiment, the piston 312 has a very aggressivecutback 314 forming the top land to enlarge the crevice volume 316between the top land and the liner. The crevice volume 316 is now solarge that it will no longer quench combustion in this area thusallowing fuel in this volume to bum resulting in reduced UHC. FIG. 49illustrates yet another embodiment including a cup or chamber 320 formedin the cylinder head 322 of the engine. The fuel injector 324 ispositioned to inject fuel directly into cup 320 early in the compressionstroke. Because air is pushed into cup 320, the fuel does not exit thecup. After compression ignition occurs, the products can pass throughthe relatively large passage or throat 326 between the cup 320 and themain cylinder 328. The fuel is well mixed because of the turbulence ofthe air entering the cup. Because there are no crevices in the cup andbecause the fuel does not leave the cup until after combustion iscompleted, UHC are extremely low. The cup could easily be coated with athermal barrier coating to reduce heat losses.

[0208]FIGS. 50a and 50 b illustrate a cup design for a four strokeengine. The exhaust and intake valves 330 are arranged around a cup 332in the head 334. Cup 332 may be positioned directly above the combustionchamber 336 as shown in FIG. 50a or offset to allow more room for thevalves 330 as shown in FIG. 50b. Another possibility is to include asmall auxiliary valve in the cup to allow the products to exit the cupmore efficiently. This valve could open after the main exhaust valveopens so that the auxiliary exhaust valve in the cup would not openagainst a high pressure. In this case, the auxiliary exhaust valve couldbe electronically operated. The timing of opening and closing of thisvalve could be used to vary the residual mass fraction which would allowcontrol over the SOC using this auxiliary valve. Also, an opposed pistonengine, as discussed hereinabove, may be used to substantially reducethe crevice volume by avoiding a cylinder head and the associatedcrevices.

[0209] Now referring to FIG. 1a, another embodiment of the presentinvention for reducing emissions is disclosed. Specifically, thisembodiment controls UHC and CO by heating the upper portion of thecylinder liner 49 to cause oxidation of the charge in the crevices. Aheater 51 is incorporated into the upper part of the liner. The heatercould be any type of heater capable of effectively producing heat, suchas an electrical resistance heater. The heater heats the gas in thecrevice above the top ring when the piston nears TDC. This heating willcause the gas to be less dense resulting in a smaller mass of chargeremaining in the crevice. The charge leaving the crevice will be at ahigher temperature due to the heating thus increasing the tendency ofthe charge to react and form CO₂ instead of CO and UHC.

[0210] Also, a glow plug may be used to heat the combustion gases toreduce emissions by enabling a larger portion of the crevice volume tobe burned. It has been determined by applicants that a glow plug willhave only a slight impact on the SOC. Since SOC changes only slightlywhen the glow plug is turned on, it does not appear that the glow plugis initiating combustion. It is more likely that when the glow plug,which was located in a spacer plate, is turned on, it gradually warms upthe gas in the crevice volume. This increase in temperature issufficient to speed up the onset of rapid combustion, and burn more ofthe fuel than would have been burned without the glow plug on, resultingin a slight increase in GIMEP.

[0211] The present engine, as shown in FIGS. 1a and 1 b, may also beoperated as a multi-mode engine which changes modes of operation basedon the operating conditions or needs of the particular application. Forinstance, the engine may be operated on diesel fuel only as aconventional diesel engine, as a modified diesel engine with diesel fuelbeing injected earlier in the compression event than the conventionaldiesel engine, as a spark-ignited engine using spark plug 56 (FIG. 1a)or as a PCCI engine. This type of spark ignited/compression ignited DI(direct injection) variable compression ratio engine provides acombination of low emissions, high power density, and ease of starting.

[0212] This engine operates in the following different modes dependingon the current operating conditions/needs of the engine.

[0213] 1) Medium compression ratio (˜10:1), early injection (fuelinjected during intake stroke or very early in the compression stroke)nearly homogeneous:

[0214] a) Overall lean mixture, spark ignited —allows low NOx, highbrake mean effective pressure (BMEP) operation, as well as medium BMEPoperation.

[0215] b) Stoichiometric mixture, spark ignited—allows high NOx highBMEP transient operation, as well as low NOx operation with a 3 waycatalyst.

[0216] 2) High compression ratio (˜15:1), early injection, nearlyhomogeneous, very lean ((φ<0.5), compression ignition - allows very lowNOx medium BMEP and low BMEP operation.

[0217] 3) High compression ratio (˜15:1), late injection, stratifiedcharge:

[0218] a) Spark ignited —allows medium NOx, medium BMEP unthrottledoperation, and low BMEP operation.

[0219] b) Compression ignition —allows medium NOx medium and low BMEPunthrottled operation.

[0220] 4) Low compression ratio (˜8:1), early injection, nearlyhomogeneous, spark ignited:

[0221] a) Lean burn —allows very high BMEP operation.

[0222] b) Stoichiometric —allows very high BMEP operation.

[0223] 5) Medium compression ratio (˜10:1), late injection, stratifiedcharge, spark ignited —allows medium NOx, medium and low BMEP, and highBMEP operation.

[0224] 6) Very high compression ratio (˜20:1), lean bum, earlyinjection, nearly homogeneous, compression ignition —allows the engineto be started in PCCI mode.

[0225] The key here is to take full advantage of the variablecompression ratio. Starting of the engine can be achieved with sparkignition at a lower compression ratio and then transitioning to highcompression ratio, lean PCCI operation for low NOx. For less severe(e.g. not as cold) conditions, engine starting could be achieveddirectly with very high compression ratio PCCI operation. At low andmedium loads, the engine can operate in a PCCI mode as the compressionratio is adjusted to keep the start of combustion near the optimum crankangle. For high load requirements, the air/fuel ratio can be enriched,compression ratio lowered, and the engine can be spark ignited. In orderto handle sudden transients, the engine may go into one of the lateinjection modes where richer air/fuel ratios are possible without enginedamage.

[0226] In the multi-mode engine, ECU 20 (FIG. 1) functions with acontrol strategy for controlling the various controlled features of theengine to effectively switch between, and operate in, the differentmodes in order to achieve a variety of objectives. For example, themulti-mode engine achieves low NOx emissions in the PCCI mode whileenhancing startability by providing a high compression ratio or sparkignition. In addition, the engine can achieve a high cylinder pressureat high BMEP by switching to a lower compression ratio spark-ignitedmode. The multi-mode engine also permits stable combustion to occurafter switching to late injection which results in a stratified chargeby rapidly adjusting the compression ratio. Also, fuel consumption canbe effectively controlled using high compression, PCCI operation andstratified charge operation requiring no throttling which have excellentthermal efficiency. This operation also improves transient response bygoing from PCCI to late injection, stratified charge to suddenly enrichthe mixture. This multi-mode engine can also effectively minimize knock,and therefore knock damage, by effectively operating at lean PCCI orstratified charge or low compression ratio, lean burn or stoichiometricconditions. Of course, the engine operates to effectively control thestart of combustion during PCCI operation by varying, for example, asdiscussed hereinabove, the temperature and/or the equivalence ratioand/or the pressure and/or the air/fuel mixture autoignition properties.This engine could run on a variety of fuels like gasoline or dieselfuel.

[0227] Another operating mode is dual injection in which an earlyinjection is used to create a lean charge for PCCI operation. A second,late injection then adds a small amount of stratified fuel which can beeither spark or compression ignited to help ignite the remaining fuel.This mode is similar to diesel pilot operation but would only be usedduring transition between the different modes of operation or duringengine starting. Applicants have studied the effects of diesel pilotoperation on emissions. FIG. 51 shows a comparison of the normalizedheat release rate versus crank angle for the three different dieselpilot injection quantities into a PCCI engine operating on propane. Amicro-pilot injection of 0.1% resulted in good heat release placementwith no measurable increase in FSNOx. A diesel pilot of an amountestimated to supply 3.6% of the fuel energy resulted in a heat releasecurve having substantially the same shape as the previous case. The SOCis slightly more advanced than that of the 0.1% case despite a lower IMTand constant equivalence ratio. Also, FSNOx emissions have increasedover the 0.1% case from zero to 3.9 g/kg. The final curves illustratesthe heat release for a case with ˜18 % of the fuel energy coming fromthe diesel pilot. The heat release rate curve is shaped the same as theclassic diesel heat release rate curve with a premixed burn spike and adiffusion burn region. Also, the FSNOx (15.3 g/kg) and FSHC (478 g/kg)are significantly higher than in the cases with smaller diesel pilots.

[0228] With respect to diesel pilot injection, as the percentage of fuelenergy from the pilot increases, the start of combustion (SOC) becomesmore advanced, despite the lowering of IMT and a constant equivalenceratio. This earlier SOC is caused by the diesel fuel autoignitingearlier than the propane. As the percentage of pilot increases, the heatreleased by the pilot during the compression stroke increases, leadingto higher temperatures earlier in the cycle. Higher temperaturesincrease the chemical reaction rates of reactions involving propane,leading to earlier autoignition of the propane. Therefore, extremely lowNOx levels and good heat release placement can be achieved when using avery small diesel pilot or micropilot, preferably less than 4% of thetotal fuel energy.

[0229] Applicants have also studied the control of noise associated withPCCI combustion. Level of noise generated by PCCI combustion is relatedto the knock intensity. Thus, as knock intensity is decreased, noisesdecreases. As shown in FIGS. 4a, 4 c and 6, lowering cylinder pressure,for example, by retarding the SOC, substantially decreases the knockintensity and, therefore, noise. The present engine and control systempermits continuous PCCI combustion with minimal noise by avoidingexcessive peak cylinder pressures while maintaining the requiredcylinder pressure necessary for efficient, low emission PCCI combustionand the desired power output.

[0230] The control system of the present invention operates to activelyand variably control the mixture's temperature, pressure, autoignitioncharacteristic and equivalence ratio to ensure that the combustion eventoccurs between 20 crank angle degrees BTDC and 35 crank angle degreesATDC. The control system achieves this function by using combustionsensor 16, e.g. pressure sensor, to signal the start of combustion orthe location of the heat release event for each cycle. Also, ECU 20,which receives the signals from sensor 16, determines whether the SOC isoccurring within a predetermined crank angle range and determineswhether the duration of combustion is within a predetermined desiredcrank angle range. One conventional way for the ECU to determine theoptimum SOC would be to use a look-up table. If the SOC and/or theduration of combustion are outside the predetermined crank angle range,then ECU 20 determines the appropriate control variable or variables toadjust, and generates and sends the appropriate signal 22 to the chosencontrol mechanism or mechanisms, e.g. air cooler 28, heater 30 glow plug44, fuel control valves 39, 41, variable compression ratio device 38,etc., as discussed hereinabove. The control variables are varied asrequired to maintain the timing of the start of PCCI combustionpreferably between 20 crank angle degrees BTDC and 10 crank angledegrees ATDC, and to maintain the duration of combustion in the range of5-30 crank angle degrees.

[0231] Applicants have determined that, in order to initiate andmaintain PCCI combustion upon start-up in a cold engine, the conditionsin the cylinders, e.g. temperature and/or pressure, must be activelyinfluenced. For example, the intake air temperature could be raisedusing heater 30 and/or a glow plug 44, and/or the in-cylinder wallsheated using a cylinder wall heater 51 and/or an enginecoolant/lubricating oil heater. Also, the in-cylinder pressure andtemperature could be increased using variable compression ratio device38. Another effective control feature for enhancing startability is toadd small amounts of ozone to the intake air supply using injector 42,or into the cylinder using injector 40. Alternatively, or additionally,one of the fuel supplies could have a high autoignition property, e.g.low octane number. Also, the engine may be operated in a non-PCCI, forexample, as a spark-ignition, dual fuel or diesel engine, duringstarting of the engine. One or a combination of these controls arevaried, in accordance with the principles discussed hereinabove withrespect to each control feature, to cause PCCI combustion to occur. Asthe engine starts, the ECU will monitor the start of combustion andduration of combustion by receiving combustion data, e.g. pressuresignals, from sensor 16 throughout engine operation.

[0232] Once the engine is warmed up, the SOC and duration of combustionwill vary due to the sensitivity of PCCI combustion to the temperatureand pressure history. Small variations in the numerous factors affectingtemperature and pressure history, such as combustion chamber walltemperature, IMT, equivalence ratio, IMP, etc. result in significantvariation in the SOC and the duration of combustion. During operation,the control system of the present invention will vary one or more of thecontrol variables, that is, temperature, pressure, air/fuel mixtureautoignition properties and/or equivalence ratio, using the variouscontrol mechanisms discussed hereinabove, in such a manner to maintainthe SOC and duration of combustion in the desired ranges. For example,applicants have shown that SOC can be advanced from 50° ATDC to 0.50 °BTDC by increasing the IMT from 184° F. to 195° F., as shown in FIG. 8.Applicants have also shown that increasing CR, which raises thein-cylinder temperatures, can be used to advance SOC. For example, FIG.21 shows that increasing CR from 14:1 to 22:1 advanced the SOC from 2°ATDC to 13° BTDC when the equivalence ratio was 0.35 and IMT was 380 K.In addition, applicants have shown that increasing RMF to raise thetemperature of the charge also can be used to advance SOC. When RMF wasincreased by adjusting exhaust valve lash from 0.025“to 0.046”, the SOCadvanced from 6.40 ATDC to 1.7° ATDC, as shown in FIG. 16. Heat transferto the charge, whether from active heating elements or hot surfaces suchas the combustion chamber walls, has also been shown to advance SOC.Applicants have also shown that, with a glow plug installed in thecombustion chamber, the SOC retarded from 0.60° ATDC to 1.50° ATDC afterthe glow plug had been turned off, as shown in FIG. 11. Applicants havedetermined, as shown in FIG. 9, that increasing combustion chamber walltemperatures from 400K to 933K can advance the SOC from 7° ATDC to 14°BTDC.

[0233] With respect to pressure control, increasing IMP serves toadvance the SOC. FIG. 31c, for example, shows that increasing IMP on thesingle cylinder engine from 52 psia to 57 psia caused the SOC to advancefrom 3.7° ATDC to 1.50° BTDC. Any method of affecting cylinder pressure,such as varying compression ratio or changing valve timing, bothillustrated above, can be used to control SOC.

[0234] With respect to equivalence ratio, applicants have determined, asshown in FIG. 38, show that increasing equivalence ratio from 0.30 to0.33 by increasing fuel flow to the engine advanced the SOC from 5.5°ATDC to 2.0° ATDC. Also, varying the autoignition properties of theair/fuel mixture by the addition of reactive species or even diluent canaffect SOC. Applicants have shown that for the case shown in FIG. 33,increasing the amount of ozone added to the charge from 0 to 36 g/kg offuel had the effect of advancing SOC from 1° ATDC to 12.5° BTDC. In onestudy where diesel fuel was used in a pilot injection to initiate SOC inan air-propane mixture, the amount of pilot used affected SOC. Forexample, when pilot quantity was increased from approximately 0.1% to18% of the total fuel energy, the SOC advanced from 2° ATDC to 10° BTDC.In one study, EGR was used as a diluent to retard SOC while holding IMTconstant with an aftercooler. As shown in FIG. 17, when EGR rate wasincreased from 2.9% to 8.0%, the SOC retarded from 1.2° ATDC to 4.20ATDC. Applicants have shown that increasing the air/fuel mixture'sresistance to autoignition by increasing octane number, for example, canbe used to retard SOC. Also, applicants have shown that when octanenumber was increased from 80 to 100, the SOC retarded from 140 BTDC to7° BTDC for a case where IMT plus reheat was 311K.

[0235] Of course, any of these control variables could be adjusted inthe opposite direction from the above examples to achieve the oppositeeffect on SOC if necessary. For example, rather than increasing IMT toadvance SOC, IMT could be decreased to retard SOC. Also, the magnitudesof such variations would be increased or decreased as necessary tomaintain the desired SOC.

[0236] Applicants have shown that the combustion or heat releaseduration can be affected by varying different parameters. As SOC isretarded, the heat release duration increases. For example, FIG. 8 showsthat as SOC is retarded, by reducing IMT from 195 degrees F. to 184degrees F., the duration increases from approximately 6 degrees toapproximately 24 degrees. Similarly, increasing the equivalence ratiodecreases the heat release duration. Applicants also believe thatincreasing the degree of temperature and equivalence ratiostratification of the charge increases the heat release duration.However, given the difficulty of measuring the degree of temperature orequivalence ratio stratification more work is needed to quantify thelevel of stratification.

[0237] Of course, given the relationship between SOC and duration, anycontrol strategy that retards SOC should also increase the duration. Bymaintaining the SOC and the duration of combustion in the desired rangeswhile controlling the equivalence ratio to ensure lean burn conditions,the control system minimizes NOx emissions. Also, the present enginedesign, also reduces UHC and CO emissions by minimizing the crevices inthe cylinder thereby minimizing the unburned gases as shown in FIGS.48a-50 b.

[0238] During operation, balancing the combustion processes between thecylinders of the engine of FIG. 1b can be accomplished by varying any ofthe control variables used to control the SOC, as discussed hereinabove.The ECU 20 compares the SOC and duration of combustion data provided bysensor 16 for each cylinder. When the data indicates that the SOC and/orduration of combustion of one or more cylinders is occurring outside apredetermined crank angle range, the ECU will determine the appropriatecontrol variable or variables most effective for the given operatingconditions and generates a control signal for controlling the controlvariable to cause the SOC and/or duration of combustion to adjust so asto fall within the desired range. Applicants have determined thatcylinder balancing is best achieved by controlling equivalence ratio,adding ozone to the mixture, controlling individual heaters associatedwith each cylinder intake port, varying compression ratio using device38 or variable valve timing, adding oil via pilot injection or port fuelinjection, port injection of water and/or any of the methods discussedhereinabove for varying EGR or RMF. Any of these or other forms ofcombustion control could be used alone, or in a variety of combinations,to enhance combustion balancing control. For example, the combustioncontrol provided by the multiple fuel/additive system describedhereinabove could be enhanced by providing variable valve timing and/orcombustion chamber surface temperature cooling, e.g. engine coolant, orpiston cooling nozzle control. Also, one or more glow plugs 44 (FIG. 1a)may be used as an inexpensive, easy method of achieving at least partialcontrol over combustion balancing between the cylinders. It may also bepossible to control the EGR rate for each cylinder in order to balancecombustion quality.

INDUSTRIAL APPLICABILITY

[0239] The present PCCI engine and control system may be used in anystationary or nonstationary power plant, including any automotive,industrial, marine or military application. The present PCCI engine andcontrol system is especially advantageous in any power generationapplication where low emissions are desirable.

We claim:
 1. A premixed charge compression ignition internal combustionengine, comprising: an engine body; a combustion chamber formed in theengine body; an intake air system for delivering intake air, includingat least one of air and a mixture of air and fuel, to said combustionchamber; combustion history control system for controlling a combustionhistory of future combustion events to reduce emissions and optimizeefficiency, said combustion history control system including a mixtureautoignition property control system for varying an autoignitionproperty of the mixture, said mixture autoignition property controlsystem including a first fuel supply for supplying a first fuel to theengine and a second fuel supply connected to at least one of said intakeair system and said combustion chamber for supplying a second fuel tothe engine, said first fuel having a first autoignition property andsaid second fuel having a second autoignition property different fromsaid first autoignition property; and a processor adapted to control anamount of said second fuel delivered to said at least one of said intakeair system and said combustion chamber to optimize engine operation.